JP2008082216A - Compression expansion turbine system - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a compression expansion turbine system capable of preventing grease in a rolling bearing from being scattered by flow-in air due to pressure difference, and improving long term durability of the rolling bearing. <P>SOLUTION: In this compression expansion turbine system, rolling bearings 15, 16 and a magnetic bearing are used for supporting a main shaft 13 in parallel. An electromagnet 17 of the magnetic bearing is attached to a spindle housing 14 to oppose to a thrust plate 13a provided on the main shaft without contacting the same. A compressor side impeller 6a and a turbine side impeller 7a are fitted to the thrust plate 13a and the common main shaft 13, and the compressor side impeller 6a is driven by power generated by the turbine side impeller 7a. A sleeve 21A forming a gap A to be a non-contact seal is provided with adjoining a high pressure side of the rolling bearing 15 in the compressor side impeller 6a. The gap A and an end part of a bearing space of the rolling bearing 15 are relatively positioned in a range where the same don't overlap in a radial direction. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a compression / expansion turbine system of an air cycle refrigeration cooling system in which air is used as a refrigerant and is used for a refrigeration warehouse, a low-temperature room below zero degrees, air conditioning, and the like.

  Since the air cycle refrigeration cooling system uses air as a refrigerant, energy efficiency is insufficient as compared with the case of using chlorofluorocarbon, ammonia gas, or the like, but it is preferable in terms of environmental protection. In addition, in facilities where refrigerant air can be directly blown into, such as a refrigerated warehouse, the total cost may be reduced by omitting the internal fan and defrost, etc. In such applications, the air cycle refrigeration cooling system Has been proposed (for example, Patent Document 1).

Further, it is known that the theoretical efficiency of air cooling is equal to or higher than that of Freon or ammonia gas in a deep coal region of -30 ° C to -60 ° C. However, it is also stated that obtaining the theoretical efficiency of the air cooling is not possible until there is an optimally designed peripheral device. The peripheral device is a compressor, an expansion turbine, or the like.
As the compressor and the expansion turbine, a turbine unit in which a compressor impeller and an expansion turbine impeller are attached to a common main shaft is used (Patent Document 1).

In addition, as a turbine compressor which processes process gas, a turbine impeller is attached to one end of the main shaft, a compressor impeller is attached to the other end, and the main shaft is supported by a journal and a thrust bearing that is controlled by an electromagnet current. A compressor has been proposed (Patent Document 2).
In addition, although it is a proposal for a gas turbine engine, in order to avoid the thrust load acting on the rolling bearing for supporting the main shaft from shortening the bearing life, the thrust load acting on the rolling bearing should be reduced by the thrust magnetic bearing. Has been proposed (Patent Document 3).
Japanese Patent No. 2623202 Japanese Patent Application Laid-Open No. 7-91760 JP-A-8-261237

As described above, as the air cycle refrigeration cooling system, in order to obtain the theoretical efficiency of air cooling that is highly efficient in the deep coal region, an optimally designed compressor and expansion turbine are required.
As the compressor and the expansion turbine, a turbine unit in which the compressor wheel and the expansion turbine wheel are attached to a common main shaft as described above is used. In this turbine unit, the compressor impeller can be driven by the power generated by the expansion turbine, thereby improving the efficiency of the air cycle refrigerator.

However, in order to obtain practical efficiency, it is necessary to keep the gap between each impeller and the housing minute. The fluctuation of the gap hinders stable high-speed rotation and causes a decrease in efficiency.
In addition, a thrust force acts on the main shaft by the air acting on the compressor impeller and the turbine impeller, and a thrust load is applied to the bearing that supports the main shaft. The rotation speed of the main shaft of the turbine unit in the air cycle refrigeration cooling system is 80,000 to 100,000 rotations per minute, which is very high compared with a bearing for general use. For this reason, the thrust load as described above causes a decrease in long-term durability and life of the bearing supporting the main shaft, and decreases the reliability of the turbine unit for air cycle refrigeration cooling. Unless such a problem of long-term durability of the bearing is solved, it is difficult to put the air cycle refrigeration cooling turbine unit into practical use. However, the technique disclosed in Patent Document 1 has not yet been solved for the deterioration of the long-term durability of the bearing against the load of the thrust load under the high-speed rotation.

  In the case where the main shaft is supported by a journal bearing made of a magnetic bearing and a thrust bearing, such as the magnetic bearing type turbine compressor of Patent Document 2, the journal bearing does not have a restriction function in the axial direction. Therefore, if there is an unstable factor in controlling the thrust bearing, it is difficult to perform stable high-speed rotation while maintaining a minute gap between the impeller and the diffuser. In the case of a magnetic bearing, there is also a problem of contact when the power is stopped.

Accordingly, the inventors of the present invention have proposed a compression / expansion turbine unit for air cycle refrigeration cooling having a configuration as shown in FIG. 8 as a means for solving the above problems (Japanese Patent Application No. 2005-239464). In this turbine unit, a roller wheel 65a for the compressor 56 and a turbine wheel 57a for the expansion turbine 57 are attached to both ends of the main shaft 63, and grease lubricated rolling bearings 65 and 66 are used as bearings for rotatably supporting the main shaft 63. In combination with an electromagnet 67 constituting a magnetic bearing, the rolling bearings 65 and 66 support a radial load, the magnetic bearing supports an axial load, and the electromagnet 67 is a thrust plate 63a provided perpendicularly and coaxially to the main shaft 63. The electromagnet 67 is controlled by the controller 69 in accordance with the output of the sensor 68 that detects the force in the axial direction.
In addition, a portion in the compressor side casing 56b that constitutes the compressor 56 with the compressor impeller 56a and a portion on the compressor impeller 56a side with respect to the rolling bearing 65, and a rolling in the turbine side casing 57b that constitutes the expansion turbine 57 with the turbine impeller 57a. The portions closer to the turbine impeller 57a than the bearing 66 are formed with inner diameter surfaces close to the main shaft 63, and non-contact seals 71 and 72 are formed on these inner diameter surfaces. The non-contact seals 71 and 72 are labyrinth seals or thread groove seals formed by arranging a plurality of circumferential grooves in the axial direction on the inner diameter surfaces of the casings 56b and 57b.

  However, in this compression / expansion turbine system, since the compressor impeller 56a and the turbine impeller 57a are attached to the common main shaft 63, a large pressure difference is generated on both sides of the rolling bearings 65 and 66 that support the main shaft 63. In the rolling bearing 65 on the compressor impeller 56a side, the compressor impeller 56a side is the high pressure side. For this reason, there is a problem that the air flowing through the non-contact seal 71 flows directly into the bearing space of the rolling bearing 65 and the grease in the rolling bearing 65 is scattered by this air flow. Further, since the high-pressure air from the compressor wheel 56a is introduced into the expansion turbine wheel 57a, the air flowing through the non-contact seal 72 flows directly into the bearing space of the rolling bearing 66, and this air flow causes the inside of the rolling bearing 66 to flow. There is a problem that the grease is scattered. Therefore, the long-term durability of the rolling bearing 65 and the rolling bearing 66 is also reduced.

An object of the present invention is to provide a compression / expansion turbine system capable of preventing the grease in the rolling bearing from being scattered by the inflow air due to the pressure difference and improving the long-term durability of the rolling bearing.
Another object of the present invention is to provide a high-precision support in the axial direction and to ensure a long-term durability of the rolling bearing, and to stably rotate at high speed while maintaining a minute gap between the diffusers of the impeller. Is to be able to do.

The compression / expansion turbine system according to the present invention supports a compressor side impeller and a main shaft that supports the turbine side impeller by using both a rolling bearing and a magnetic bearing installed in a turbine unit housing, and the compressor side impeller with respect to the rolling bearing. A compression-expansion turbine system in which a non-contact seal is provided between a turbine unit housing and a main shaft adjacent to the high pressure side of the pressure around the main shaft where a pressure difference occurs between the vehicle and the turbine side impeller, wherein the non-contact seal The radial positional relationship between the gap and the end of the bearing space formed between the inner and outer rings of the rolling bearing adjacent to the non-contact seal is a positional relationship in which no overlapping range occurs.
According to this configuration, the gap between the non-contact seal adjacent to the high-pressure side of the rolling bearing and the radial positional relationship between the end of the bearing space of the rolling bearing adjacent to the non-contact seal do not overlap each other. Because of the positional relationship, the air that has passed through the non-contact seal once hits the width surface of the inner ring or outer ring of the rolling bearing and then flows into the rolling bearing. That is, the inflow path of air from the non-contact seal to the bearing space becomes a curved path. Therefore, the speed of the air flow flowing into the bearing space is reduced, and the scattering of grease is alleviated. Thereby, the long-term durability of the rolling bearing is maintained. In addition, since this compression / expansion turbine system supports the main shaft by using both a rolling bearing and a magnetic bearing, damage when the power supply is stopped when only the magnetic bearing is supported is also avoided.

In the present invention, of the circumferential surfaces on the small diameter side and the large diameter side facing each other that constitute the gap of the non-contact seal, the inner diameter of the circumferential surface on the large diameter side is adjacent to the non-contact seal in the inner ring of the rolling bearing. It is good also as a thing smaller than the outer diameter of the end surface to do.
Of the peripheral surfaces of the small diameter side and the large diameter side facing each other that constitute the gap of the non-contact seal, the outer diameter of the peripheral surface on the small diameter side is an end surface adjacent to the non-contact seal in the outer ring of the rolling bearing. It is good also as a thing larger than the internal diameter of.
By adopting any one of the configurations described above, the radial positional relationship between the non-contact seal and the bearing space can be a positional relationship in which no overlapping range occurs. When the inner diameter of the peripheral surface on the large-diameter side of the non-contact seal is made smaller than the outer diameter of the inner ring, it is easy to design dimensions when the inner ring spacer is not provided adjacent to the high-pressure side of the rolling bearing. It becomes. When the outer diameter of the peripheral surface on the small diameter side of the non-contact seal is made larger than the inner diameter of the outer ring, an inner ring spacer is arranged adjacent to the high-pressure side of the rolling bearing, and the outer surface of the inner ring spacer is non-contact sealed. In the case of a specification for forming the gap, it is easy to design dimensions.

In the case of any one of the configurations according to the present invention, the rolling bearing supports a radial load, the magnetic bearing supports one or both of an axial load and a bearing preload, and the electromagnet of the magnetic bearing is attached to the main shaft. The compressor-side impeller is driven by the power generated by the turbine-side impeller so as to be opposed to the provided flange-shaped thrust plate made of a ferromagnetic material in a non-contact manner. May be.
In this configuration, a rolling bearing and a magnetic bearing are used together, the rolling bearing supports a radial load, and the magnetic bearing supports one or both of an axial load and a bearing preload. It is possible to provide a good support, and it is possible to ensure the long-term durability of the rolling bearing. Therefore, high-speed rotation can be stably performed while maintaining a minute gap between the diffusers of the impeller.

In this case, it has a controller that controls the electromagnet according to the output of the sensor that detects the axial force of the main shaft, and the stiffness value of the composite spring formed by the rolling bearing and the supporting system of the rolling bearing However, it is good also as what has the relationship larger than the negative rigidity value of the said electromagnet.
When the rigidity value relationship is set in this way, the phase of the mechanical system can be prevented from being delayed by 180 ° in the control band, the controlled object can be stabilized, and the controller circuit configuration can be proportional or proportionally integrated. Even with a simple configuration such as the above, stable control can be performed.

In the present invention, the magnetic bearing and the motor may be integrated as follows. That is, the rolling bearing supports a radial load, the magnetic bearing supports one or both of an axial load and a bearing preload, and the electromagnet of the magnetic bearing has a flange-like shape made of a ferromagnetic material provided on the main shaft. It is attached to the turbine unit housing so as to face the thrust plate in a non-contact manner, and a motor rotor is provided on the thrust plate, and a motor stator is arranged in the turbine unit housing so as to face the motor rotor. The compressor side impeller may be driven by either one or both of the following power and the power generated by the turbine side impeller.
Also in this configuration, since the rolling bearing and the magnetic bearing are used together, the rolling bearing supports the radial load, and the magnetic bearing supports one or both of the axial load and the bearing preload. Highly accurate support can be achieved, and long-term durability of the rolling bearing can be secured. Therefore, high-speed rotation can be stably performed while maintaining a minute gap between the diffusers of the impeller. In this configuration, since the motor rotor is provided on the thrust plate of the magnetic bearing, it is possible to reduce the size of the main shaft by shortening the spindle size by using both the magnetic bearing and the motor. Further, it is suitable for high-speed rotation.

When the magnetic bearing and the motor are integrated in this way, the controller has a controller for controlling the electromagnet according to the output of the sensor that detects the axial force of the spindle, and the rolling bearing and the support system for the rolling bearing It is good also as what has the relationship whose rigidity value of the synthetic | combination spring formed by these is larger than the negative rigidity value of the synthetic | combination spring formed with the said electromagnet and a motor.
Also in this case, in the control band, the phase of the mechanical system can be prevented from being delayed by 180 °, the controlled object can be stable, and the circuit configuration of the controller can be a simple configuration such as proportional or proportional integration. However, stable control can be performed.

In this invention, the compression / expansion turbine system may compress the inflow air by a compressor of the turbine unit, cooling by another heat exchanger, adiabatic expansion by the expansion turbine of the turbine unit, compression by pre-compression means, heat It may be used in an air cycle refrigeration cooling system that sequentially performs cooling by an exchanger, compression by a compressor of a turbine unit, cooling by another heat exchanger, and adiabatic expansion of the turbine unit by an expansion turbine.
When the compression / expansion turbine system is applied to such an air cycle refrigeration / cooling system, in the compression / expansion turbine system, a stable high-speed rotation of the main shaft can be obtained while maintaining an appropriate gap between the impellers, and the bearing has a long-term durability. Therefore, the reliability is improved as a whole of the compression / expansion turbine system and, as a whole, the air cycle refrigeration cooling system. In addition, stable high-speed rotation, long-term durability, and reliability of the main shaft bearing of the compression / expansion turbine system, which is the bottleneck of the air cycle refrigeration cooling system, improve the practical use of the air cycle refrigeration cooling system. .

The compression / expansion turbine system according to the present invention supports a compressor side impeller and a main shaft that supports the turbine side impeller by using both a rolling bearing and a magnetic bearing installed in a turbine unit housing, and the compressor side impeller with respect to the rolling bearing. A compression-expansion turbine system in which a non-contact seal is provided between a turbine unit housing and a main shaft adjacent to the high pressure side of the air pressure around the main shaft where a pressure difference occurs between the vehicle and the turbine side impeller. The radial positional relationship between the clearance of the bearing and the end of the bearing space formed between the inner and outer rings of the rolling bearing adjacent to the non-contact seal is a positional relationship that does not overlap each other. Thus, it is possible to prevent the grease in the rolling bearing from being scattered and to improve the long-term durability of the rolling bearing.
In particular, the rolling bearing supports a radial load, the magnetic bearing supports one or both of an axial load and a bearing preload, and the electromagnet of the magnetic bearing has a flange shape made of a ferromagnetic material provided on the main shaft. If it is mounted on the turbine unit housing so as to face the thrust plate without contact, it can be supported with high precision in the axial direction, and the long-term durability of the rolling bearing can be secured. High-speed rotation can be performed stably while maintaining a minute gap between the diffusers of the impeller.

  A first embodiment of the present invention will be described with reference to FIGS. FIG. 1 shows a cross-sectional view of a compression / expansion turbine unit 5 constituting the compression / expansion turbine system of this embodiment. The compression / expansion turbine unit 5 includes a compressor 6 and an expansion turbine 7. A compressor impeller 6 a of the compressor 6 and a turbine impeller 7 a of the expansion turbine 7 are respectively attached to both ends of the main shaft 13. Further, the compressor impeller 6a is driven by the power generated in the turbine impeller 7a, and no other drive source is provided.

The housing 12 of the compression / expansion turbine unit 5 includes a compressor side casing 6b, a spindle housing 14, a rolling bearing support 20 interposed between the compressor side casing 6b and the spindle housing 14, and a turbine side casing 7b. Configured.
In FIG. 1, a compressor 6 is composed of a compressor impeller 6a and the compressor casing 6b opposed to the compressor impeller 6a via a minute gap d1, and sucked in the axial direction from a suction port 6c at the center. The air is compressed by the compressor wheel 6a and discharged from the outlet (not shown) at the outer peripheral portion as indicated by the arrow 6d.
The expansion turbine 7 is composed of a turbine impeller 7a and a turbine casing 7b opposed to the turbine impeller 7a via a minute gap d2, and the air sucked from the outer periphery as indicated by an arrow 7c The vehicle 7a adiabatically expands and discharges in the axial direction from the central outlet 7d.

  In this compression / expansion turbine unit 5, the main shaft 13 is supported by a plurality of rolling bearings 15 and 16 in the radial direction, and either or both of the axial load and the bearing preload applied to the main shaft 13 are supported by an electromagnet 17 that is a magnetic bearing. It is supposed to be. The compression / expansion turbine unit 5 includes a sensor 18 that detects an axial load acting on the rolling bearing 16, and a magnetic bearing controller 19 that controls the supporting force of the electromagnet 17 according to the output of the sensor 18. is doing. The electromagnet 17 is installed in the spindle housing 14 so as to face the both surfaces of a flange-like thrust plate 13a made of a ferromagnetic material that is provided as an integral structure perpendicularly and coaxially with the main shaft 13 in the center of the main shaft 13. The thrust plate 13a is an electromagnet target. The material of the main shaft 13 is low carbon steel with good magnetic properties.

  The rolling bearings 15 and 16 that support the main shaft 13 have a function of restricting the position in the axial direction. In this embodiment, angular ball bearings are used. A deep groove ball bearing may be used instead of the angular ball bearing. In the case of a deep groove ball bearing, it has a thrust support function in both directions, and has the effect of returning the axial position of the inner and outer rings to the neutral position. These two rolling bearings 15 and 16 are disposed near the compressor impeller 6a and the turbine impeller 7a in the unit housing 12, respectively.

  The main shaft 13 is a stepped shaft having a large diameter portion 13c at the center and small diameter portions 13d at both ends. The bearings 15 and 16 on both sides have their inner rings 15a and 16a fitted into the small diameter portion 13d in a press-fit state, and one of the width surfaces engages with a stepped surface between the large diameter portion 13c and the small diameter portion 13d. Further, a through-hole (not shown) provided in the central portion of the compressor impeller 6a and the turbine impeller 7a is fitted into the small diameter portions 13d at both ends of the main shaft 13 in a press-fitted state, so that the compressor impeller 6 a and a turbine impeller 7 a are attached to the main shaft 13.

  The sensor 18 is provided on the stationary side in the vicinity of the rolling bearing 16 on the turbine impeller 7a side, that is, on the spindle housing 14 side. The rolling bearing 16 provided with the sensor 18 in the vicinity thereof has an outer ring 16 b fitted in the bearing housing 23 in a fixed state. The bearing housing 23 has an inner flange 23a that is formed in a ring shape and engages with the width surface of the outer ring 16b of the bearing 16 at one end, and is movable in the axial direction on an inner diameter surface 24 provided on the spindle housing 14. It is mated. The inner collar 23a is provided at the center side end in the axial direction.

  The sensors 18 are distributed and arranged at a plurality of circumferential locations (for example, two locations) around the main shaft 13, a width surface on the inner flange 23 a side of the bearing housing 23, and one electromagnet 17 that is a member fixed to the spindle housing 14. It is interposed between. The sensor 18 is applied with preload by a sensor preload spring 25. The sensor preload spring 25 is housed in a housing recess provided in the turbine-side casing 7 b and biases the outer ring 16 b of the rolling bearing 16 in the axial direction. The sensor preload spring 25 is moved through the outer ring 16 b and the bearing housing 23. Preload. The sensor preload spring 25 is composed of, for example, coil springs provided at a plurality of locations in the circumferential direction around the main shaft 13.

  The preload by the sensor preload spring 25 is to enable the sensor 18 that detects the thrust force by the pressing force to detect any movement of the main shaft 13 in the axial direction. The magnitude is greater than the average thrust force acting on the main shaft 13 in a normal operating state.

  The rolling bearing 15 on the non-arrangement side of the sensor 18 is installed so as to be movable in the axial direction with respect to the rolling bearing support portion 20, and is elastically supported by a bearing preload spring 26. In this example, the outer ring 15b of the rolling bearing 15 is fitted to the inner diameter surface of the rolling bearing support 20 so as to be movable in the axial direction, and the bearing preload spring 26 is interposed between the outer ring 15b and the compressor casing 6b. ing. The bearing preload spring 26 biases the outer ring 15b so as to oppose the step surface of the main shaft 13 with which the width surface of the inner ring 15a is engaged, and applies preload to the rolling bearing 15. The bearing preload spring 26 includes coil springs and the like provided at a plurality of locations in the circumferential direction around the main shaft 13 and is accommodated in the accommodating recesses provided in the compressor side casing 6b. The bearing preload spring 26 has a smaller spring constant than the sensor preload spring 25.

  Between the compressor impeller 6a and the rolling bearing 15 in the vicinity thereof, and between the turbine impeller 7a and the rolling bearing 16 in the vicinity thereof, an inner diameter surface is formed to have a diameter close to the main shaft 13, respectively. Non-contact seals comprising sleeves 21A and 21B that form seal gaps on the outer periphery are arranged. A non-contact seal is also called a gap seal.

  The sleeve 21A, which is a non-contact seal on the compressor impeller 6a side, is elastically formed on the inner diameter surface of the portion 6ba adjacent to the rolling bearing 15 in the compressor side casing 6b as shown in FIG. The support member 21B is elastically supported so as to be displaceable in the radial direction. A groove 50 is formed at the axially intermediate position of the outer diameter surface of the sleeve 21A, and a groove 51 is also formed on the inner diameter surface of the bearing adjacent portion 6ba of the compressor side casing 6b facing the groove 50. The elastic support member 21 </ b> B is engaged over 50 and 51.

  The sleeve 21 </ b> A which is a non-contact seal on the compressor impeller 6 a side is adjacent to the rolling bearing 15 on the high pressure side. The radial positional relationship between the gap A of the non-contact seal by the sleeve 21A and the end of the bearing space S formed between the inner and outer rings 15a and 15b of the rolling bearing 15 adjacent to the non-contact seal is an overlapping range. The positional relationship does not occur. Specifically, the inner diameter D1 of the inner peripheral surface of the sleeve 21A that becomes the large-diameter-side peripheral surface Ab among the small-diameter and large-diameter peripheral surfaces Aa and Ab that constitute the clearance A of the non-contact seal. However, it is made smaller than the outer diameter D2 of the end surface adjacent to the non-contact seal in the inner ring 15a of the rolling bearing 15.

  In place of the configuration of FIG. 2, as shown in FIG. 4, an inner ring spacer 41 is provided on the inner periphery of the sleeve 21A, which is a non-contact seal on the compressor impeller 6a side, and a non-contact seal gap A is configured. Of the peripheral surfaces Aa ′ and Ab ′ on the small diameter side and large diameter side facing each other, the outer diameter D3 of the peripheral surface Aa ′ on the small diameter side is set to the end surface adjacent to the non-contact seal in the outer ring 15b of the rolling bearing 15. It may be larger than the inner diameter D4. The inner ring spacer 41 is fitted to the outer diameter surface of the main shaft 13.

  In FIG. 1, the sleeve 22A, which is a non-contact seal on the turbine impeller 7a side, is also supported by the elastic support member 22B, similarly to the sleeve 21A, which is a non-contact seal on the compressor impeller 6a side. That is, it is supported by the elastic support member 22B on the inner diameter surface of the turbine side casing 7b adjacent to the rolling bearing 16 so as to be elastically movable in the radial direction. A groove is formed at the axially intermediate position of the outer diameter surface of the sleeve 22A, and a groove is also formed on the inner diameter surface of the bearing adjacent portion of the turbine-side casing 7b facing the groove. The elastic support member 22B is engaged.

The seal gap between the sleeves 21A, 22A and the main shaft 13 is preferably 15 μm or less, and if it is 10 μm or less, there is almost no air leakage in the gap. Each sleeve 21A, 22A is made of carbon or a copper alloy (including pure copper), and some contact between the main shaft 13 and the sleeves 21A, 22A is allowed.
Here, the support portions of the sleeves 21A and 22A serving as the non-contact seals are the compressor side casing 6b and the turbine side casing 7b. However, depending on the structure of the unit housing 12, the sleeve 21A on the compressor impeller 6a side may be provided. The sleeve 22A on the turbine impeller 7a side may be supported by the inner diameter surface of the rolling bearing support portion 20 and the inner diameter surface of the spindle housing 14, respectively. Further, the non-contact seal constituted by the sleeves 21 and 22A may be applied only to the compressor impeller 6a side or only to the turbine impeller 7a side.

  A gap 27 for restricting heat inflow from the compressor side casing 6b to the rolling bearing support portion 20 is formed on a surface where both end surfaces of the compressor side casing 6b and the rolling bearing support portion 20 on the compressor 6 side are coupled. . In this embodiment, a protrusion 20a is formed on the outer diameter side of the end face of the rolling bearing support portion 20, and the gap 27 is formed by surface contact or point contact of the protrusion 20a with the compressor side casing 6b. A protrusion may be formed on the end surface of the side casing 6b. The protrusions 20a may be formed in a ring shape over the entire circumference in the circumferential direction, or may be formed by being intermittently distributed in the circumferential direction.

  A cooling path 28 for cooling the unit housing 12 is formed in the vicinity of the gap 27 in the rolling bearing support portion 20. The cooling path 28 may be formed in the compressor side casing 6b. Water, oil, or air is used as a cooling medium flowing through the cooling path 28. In this way, by configuring a cooling system that allows the cooling medium to flow through the cooling path 28, the temperature of the rolling bearing support 20, and further the temperature of the rolling bearing 15, is an appropriate temperature that is lower than the allowable temperature of the rolling bearing 15 to be used. Controlled.

The dynamic model of the magnetic bearing device in the compression / expansion turbine unit 5 can be constituted by a simple spring system. That is, in this spring system, a combined spring composed of the rolling bearings 15 and 16 and a support system for these bearings (sensor preload spring 25, bearing preload spring 26, bearing housing 23, etc.) and a spring of the electromagnet 17 are arranged in parallel. This is the configuration. In this spring system, the composite spring composed of the rolling bearings 15 and 16 and the support system for these bearings has rigidity acting in proportion to the amount of displacement in the direction opposite to the displaced direction, whereas the electromagnet 17. This spring has a negative stiffness that acts in proportion to the amount of displacement in the displaced direction.
For this reason, the magnitude relationship between the stiffness of the composite spring and the negative stiffness of the spring of the electromagnet 17 is as follows.
Synthetic spring stiffness <electromagnet negative stiffness (1)
In this case, since the phase of the mechanical system is 180 ° delayed and becomes an unstable system, it is necessary to add a phase compensation circuit in advance in the magnetic bearing controller 19 that controls the electromagnet 17. It becomes complicated.

Therefore, in this embodiment, the magnitude relationship between the stiffness of the composite spring and the negative stiffness of the spring of the electromagnet 17 is as follows.
Rigidity value of synthetic spring> Negative stiffness value of electromagnet (2)
It is said. As a result, the phase of the mechanical system can be prevented from being delayed by 180 ° in the control band, so that the control target of the controller 19 can be stabilized, and the circuit configuration of the controller 19 is proportional or proportional to integral as shown in FIG. Easy to configure.
In the controller 19 of FIG. 3 shown in the block diagram, the detection outputs P1 and P2 of each sensor 18 are adjusted by the sensor output calculation circuit 31, and the calculation result is compared with the target value of the reference value setting means 33 by the comparator 32. The control signal of the electromagnet 17 is calculated by performing a proportional integration (or proportional) process that is appropriately set according to the compression / expansion turbine unit 5 by the PI compensation circuit (or P compensation circuit) 34. Is calculated. The output of the PI compensation circuit (or P compensation circuit) 34 is input to power circuits 37 and 38 that drive the electromagnets 171 and 172 in each direction via diodes 35 and 36, respectively. The electromagnets 171 and 172 are a pair of electromagnets 17 opposed to the thrust plate 13a shown in FIG. 1, and only the attractive force acts. Therefore, the direction of current is determined in advance by the diodes 35 and 36, and the two electromagnets 171 and 172 are used. Is driven selectively.

The compression / expansion turbine unit 5 having this configuration is applied to, for example, an air cycle refrigeration cooling system so that air serving as a cooling medium can be efficiently exchanged by a heat exchanger (not shown here) at a subsequent stage. Then, the air cooled by the heat exchanger and cooled by the heat exchanger at the subsequent stage is cooled and discharged by adiabatic expansion to a target temperature, for example, a very low temperature of about −30 ° C. to −60 ° C., by the expansion turbine 7. Used to be.
In such a use example, the compressor-expansion turbine unit 5 has the compressor impeller 6a and the turbine impeller 7a attached to the common main shaft 13, and the compressor impeller 6a is driven by the power generated by the turbine impeller 7a.

In this compression / expansion turbine unit 5, since the compressor impeller 6a and the turbine impeller 7a are attached to the common main shaft 13, the pressure difference between both sides of the main shaft 13 is large, and the rolling bearing on the compressor impeller 6a side which is the high pressure side. 15, air tends to flow into the rolling bearing 15 from the gap A of the non-contact seal formed by the sleeve 21 </ b> A. When the flow rate of the inflowing air is high, the grease in the rolling bearing 15 is scattered.
However, in this embodiment, the inner diameter D1 of the sleeve 21A that becomes the large-diameter peripheral surface Ab among the small-diameter and large-diameter peripheral surfaces Aa and Ab that constitute the gap A of the non-contact seal is The inner diameter 15a of the rolling bearing 15 is smaller than the outer diameter D2 of the end face adjacent to the non-contact seal. Therefore, the gap A of the non-contact seal and the end of the bearing space of the rolling bearing 15 are in a positional relationship in which no overlapping range occurs in the radial direction. Therefore, the air that has passed through the gap A of the non-contact seal once hits the width surface of the inner ring 15 a of the rolling bearing 15 and then flows into the rolling bearing 15. That is, the inflow path of air from the gap A of the non-contact seal to the bearing space becomes a curved path. Therefore, the speed of the air flow flowing into the bearing space is reduced, and the scattering of grease in the rolling bearing 15 is alleviated. Thereby, long-term durability of the rolling bearing 15 is maintained.

  In the case of the configuration of the example in FIG. 4, the air that has passed through the gap A of the non-contact seal once hits the width surface of the outer ring 15 b of the rolling bearing 15 and then flows into the rolling bearing 15. Similar to the above, the flow velocity is reduced, and the scattering of grease on the rolling bearing 15 is mitigated.

  In this embodiment, since the sleeves 21A and 22A serving as non-contact seals provided adjacent to the rolling bearings 15 and 16 on the compressor impeller 6a and turbine impeller 7a side are elastically supported, excellent sealing performance is obtained. can get. Specifically, since the sleeves 21A and 22A can move in the radial direction within the elastic deformation range of the elastic support members 21B and 22B, and the sleeves 21A and 22A and the main shaft 13 constitute a perfect circle bearing, the sleeve 21A 22A can follow the swing of the main shaft 13. By improving the sealing performance in this way, the rolling bearings 15 and 16 due to the pressure difference between the compressor 6 side and the expansion turbine 7 side, or the rolling bearings 15 and 16 and the bearing support portion (the rolling bearing support portion 20 and the spindle housing). 14) The air leakage from the contact surface with 14) can be reduced, and the efficiency reduction of the compression / expansion turbine unit 5 can be improved. Further, since it is difficult for the grease to dry and scatter in the rolling bearings 15 and 16 due to the passing air, the long-term durability of the rolling bearings 15 and 16 can be further improved.

  Further, in this embodiment, the compressor impeller 6a and the turbine impeller 7a are fitted to the main shaft 13 common to the thrust plate 13a, and the compression / expansion for driving the compressor impeller 6a by the power generated in the turbine impeller 7a. In the compression / expansion turbine unit 5 constituting the turbine system, the magnetic bearing device having the above-described configuration is applied to support the main shaft 13, so that the stable high speed of the main shaft 13 is maintained while maintaining appropriate gaps d 1 and d 2 between the impellers 6 a and 7 a. The rotation is obtained, and the long-term durability and the life of the rolling bearings 15 and 16 are improved.

  That is, in order to ensure the compression and expansion efficiency of the compression / expansion turbine unit 5, it is necessary to keep the gaps d1 and d2 between the impellers 6a and 7a and the casings 6b and 7b minute. For example, when this compression / expansion turbine unit 5 is applied to an air cycle refrigeration cooling system, ensuring this efficiency is important. On the other hand, since the main shaft 13 is supported by the rolling bearings 15 and 16, the axial position of the main shaft 13 is regulated to some extent by the restriction function of the axial position of the rolling bearing, and the impellers 6a and 7a and the housing 6b. , 7b can be kept constant.

However, a thrust force is applied to the main shaft 13 of the turbine unit 5 by the pressure of air acting on the impellers 6a and 7a. Further, the turbine unit 5 used in the air cooling system rotates at a very high speed of about 80,000 to 100,000 rotations per minute, for example. Therefore, when the thrust force acts on the rolling bearings 15 and 16 that rotatably support the main shaft 13, the long-term durability of the rolling bearings 15 and 16 is reduced.
In this embodiment, since the thrust force is supported by the electromagnet 17, the thrust force acting on the rolling bearings 15 and 16 for supporting the main shaft 13 can be reduced while suppressing an increase in torque without contact. In this case, since the sensor 18 for detecting the thrust force acting on the rolling bearing 16 and the magnetic bearing controller 19 for controlling the supporting force by the electromagnet 17 according to the output of the sensor 18 are provided, the rolling bearings 15 and 16 are provided. Can be used in an optimum state against the thrust force according to the bearing specifications.

  6 and 7 show still another embodiment of the present invention. In the compression / expansion turbine system of this embodiment, in the first embodiment shown in FIG. 1, the main shaft 13 is supported by rolling bearings 15 and 16 and an electromagnet 17 as a magnetic bearing, and the main shaft 13 is driven to rotate. The motor 39 is provided. The motor 39 is controlled by the motor controller 29 independently of the electromagnet 17.

  The electromagnet 17 is provided on each surface of two flange-shaped thrust plates 13a and 13b made of a ferromagnetic material provided as an integral structure perpendicular to and coaxially with the main shaft 13 so as to be aligned in the axial direction at the axial intermediate portion of the main shaft 13. A pair is installed in the spindle housing 14 so as to face each other without contact. Specifically, one of the electromagnets 17 constituting the magnetic bearing unit is opposed to this one surface in a non-contact manner with the one surface facing the expansion turbine 7 of the thrust plate 13a located near the expansion turbine 7 as an electromagnet target. Installed in the spindle housing 14. Further, the other electromagnet 17 constituting the magnetic bearing unit has one surface facing the compressor 6 side of the thrust plate 13b located near the compressor 6 as an electromagnet target, and faces the spindle housing 14 so as to face this one surface in a non-contact manner. Installed.

The motor 39 is a motor unit including a motor rotor 39a provided on the main shaft 13 along with the electromagnet 17, and a motor stator 39b facing the motor rotor 39a in the axial direction. Specifically, the motor rotor 39a that constitutes one part of the motor unit is arranged on each side of the main shaft 13 opposite to the side where the electromagnets 17 of the thrust plates 13a and 13b face each other at an equal pitch in the circumferential direction. By arranging the permanent magnets 39aa arranged side by side, a pair of left and right are configured. The permanent magnet 39aa is bonded and fixed to each side of the thrust plates 13a and 13b with an adhesive. Thus, between the permanent magnets 39aa arranged opposite to each other in the axial direction, the magnetic poles are set to be different from each other. Since the main shaft 13 is made of low carbon steel having good magnetic properties, the thrust plates 13a and 13b provided so as to be integrated with the main shaft 13 are also used as the back yoke and the electromagnet target of the permanent magnet 39aa. it can.
The motor stator 39b, which is another part of the motor unit, is arranged without a core so as to face each surface of both the motor rotors 39a in a non-contact manner at a central position between the left and right motor rotors 39a. The coil 39ba is installed on the spindle housing 14 via the motor housing 40. The motor 39 rotates the main shaft 13 by Lorentz force acting between the motor rotor 39a and the motor stator 39b. Thus, since this axial gap type motor 39 is a coreless motor, the negative rigidity due to the magnetic coupling between the motor rotor 39a and the motor stator 39b is zero. The motor stator 39b may have a core, and the main shaft 13 may be rotated by a magnetic force acting between the motor rotor 39a and the motor stator 39b.

The dynamic model of the motor-integrated magnetic bearing device in the compression / expansion turbine unit 5 is formed by rolling bearings 15 and 16 and a support system for these bearings (sensor preload spring 25, bearing preload spring 26, bearing housing 23, etc.). And a spring system in which a synthetic spring formed by a motor unit (the motor 39 and the electromagnet 17) is arranged in parallel. In this spring system, the combined spring formed by the rolling bearings 15 and 16 and the support system of these bearings has rigidity that acts in proportion to the amount of displacement in the direction opposite to the displaced direction, whereas the motor portion. The synthetic spring formed in (1) has a negative stiffness that acts in proportion to the amount of displacement in the displaced direction.
For this reason, the magnitude relationship between the stiffnesses of the two composite springs described above is
Rigidity value of synthetic spring by bearings etc. <Negative stiffness value of synthetic spring by motor unit ... (3)
In this case, the phase of the mechanical system is delayed by 180 ° and becomes an unstable system. Therefore, in the magnetic bearing controller 19 for controlling the electromagnet 17, it is necessary to load a phase compensation circuit in advance. It becomes complicated.

Therefore, in the motor-integrated magnetic bearing device in this embodiment, the magnitude relationship between the stiffnesses of the two combined springs is as follows.
Rigidity value of the composite spring by the bearing etc.> Negative rigidity value of the synthetic spring by the motor unit (4)
It is said. In particular, in this motor-integrated magnetic bearing device, since the axial gap motor 39 is a coreless motor as described above, the negative rigidity value acting on the motor 39 can be made zero, and the motor 39 is high. Even in the state where the load operation is performed and an excessive axial load is applied, the magnitude relationship of the above-described equation (4) can be maintained.
As a result, since the phase of the mechanical system can be prevented from being delayed by 180 ° in the control band, the control target of the magnetic bearing controller 19 can be stabilized even when the motor 39 is operated at a high load and an excessive axial load is applied. The circuit configuration of the controller 19 can be configured as a simple one using proportional or proportional integration as shown in FIG. 3 in the first embodiment.

  In the motor controller 29 shown in the block diagram of FIG. 6, the phase adjustment circuit 41 adjusts the phase of the motor drive current using the rotation angle of the motor rotor 39a as a feedback signal based on the rotation synchronization command signal. The constant rotation control is performed by supplying the motor driving current supplied from the motor driving circuit 42 to the motor stator 39b. The rotation synchronization command signal is calculated according to the output of a rotation angle detection sensor (not shown) provided in the motor rotor 39a. Other configurations in this embodiment are the same as those in the first embodiment. In this embodiment, O-rings may be used as the elastic support members 21B and 22B that support the sleeves 21A and 22A serving as non-contact seals as in the embodiment shown in FIGS.

  FIG. 7 shows the overall configuration of an air cycle refrigeration cooling system using the compression / expansion turbine unit 5 in the first embodiment, for example. The motor-integrated compression / expansion turbine unit 5 in the embodiment of FIG. It may be used. This air cycle refrigeration cooling system is a system that directly cools air in a space to be cooled 10 such as a refrigeration warehouse as a refrigerant, and circulates air from an air intake port 1a to a discharge port 1b respectively opened in the space to be cooled 10. It has path 1. In this air circulation path 1, pre-compression means 2, first heat exchanger 3, compressor 6 of air cycle refrigeration cooling turbine unit 5, second heat exchanger 8, intermediate heat exchanger 9, and said turbine unit Five expansion turbines 7 are provided in order. The intermediate heat exchanger 9 exchanges heat between the inflow air near the intake port 1a in the same air circulation path 1 and the air that has been heated by the subsequent compression and cooled. The air in the vicinity of 1a passes through the heat exchanger 9a.

  The pre-compression means 2 comprises a blower or the like and is driven by a motor 2a. The first heat exchanger 3 and the second heat exchanger 8 have heat exchangers 3a and 8a for circulating a cooling medium, respectively, and a cooling medium such as water and an air circulation path in the heat exchangers 3a and 8a. Heat exchange with 1 air. Each of the heat exchangers 3 a and 8 a is connected to the cooling tower 11 by piping, and the cooling medium whose temperature is increased by heat exchange is cooled by the cooling tower 11. Note that an air cycle refrigeration cooling system that does not include the pre-compression means 2 may be used.

  This air cycle refrigeration cooling system is a system that keeps the space to be cooled 10 at about 0 ° C. to −60 ° C., and is 1 atmosphere at about 0 ° C. to −60 ° C. from the space to be cooled 10 to the inlet 1a of the air circulation path 1. Inflow of air. Note that the numerical values of temperature and atmospheric pressure shown below are examples that serve as a guide. The air that has flowed into the intake port 1a is used by the intermediate heat exchanger 9 to cool the downstream air in the air circulation path 1 and is heated to 30 ° C. The heated air remains at 1 atm, but is compressed to 1.4 atm by the pre-compression means 2, and the temperature is raised to 70 ° C. by the compression. Since the 1st heat exchanger 3 should just cool the air of 70 degreeC which raised temperature, even if it is cold water about normal temperature, it can cool efficiently and it cools to 40 degreeC.

The air at 40 ° C. and 1.4 atm cooled by heat exchange is compressed to 1.8 atm by the compressor 6 of the turbine unit 5, and the second heat is increased to about 70 ° C. by this compression. It is cooled to 40 ° C. by the exchanger 8. The 40 ° C. air is cooled to −20 ° C. by the −30 ° C. air in the intermediate heat exchanger 9. The atmospheric pressure is maintained at 1.8 atmospheric pressure discharged from the compressor 6.
The air cooled to −20 ° C. by the intermediate heat exchanger 9 is adiabatically expanded by the expansion turbine 7 of the turbine unit 5, cooled to −55 ° C., and discharged from the outlet 1 b to the cooled space 10. This air cycle refrigeration cooling system performs such a refrigeration cycle.

  In this air cycle refrigeration cooling system, in the turbine unit 5, stable high-speed rotation of the main shaft 13 can be obtained while maintaining appropriate gaps d1 and d2 between the impellers 6a and 7a, and the long-term durability of the bearings 15 and 16 can be improved. Since the long-term durability of the bearings 15 and 16 is improved by obtaining the improvement and the improvement of the life, the reliability of the turbine unit 5 as a whole and, as a whole, the air cycle refrigeration cooling system is improved. As described above, stable high-speed rotation, long-term durability, and reliability of the main shaft bearings 15 and 16 of the turbine unit 5 that are the bottleneck of the air cycle refrigeration cooling system are improved. It becomes possible.

1 is a cross-sectional view showing a compression / expansion turbine system according to a first embodiment of the present invention. FIG. 2 is an enlarged cross-sectional view around a non-contact seal on the compressor impeller side in FIG. 1. It is a block diagram which shows an example of the controller for magnetic bearings in the compression / expansion turbine system of FIG. It is an expanded sectional view which shows the modification of the non-contact seal periphery of the compressor impeller side of FIG. It is sectional drawing of the compression expansion turbine system which concerns on other embodiment of this invention. It is a block diagram which shows an example of the controller for motors in the same compression expansion turbine system. It is a systematic diagram of an air cycle refrigeration cooling system to which the same compression / expansion turbine system is applied. It is sectional drawing of a proposal example.

Explanation of symbols

2 ... Pre-compression means 3 ... First heat exchanger 5 ... Compression / expansion turbine unit 6 ... Compressor 6a ... Compressor impeller 6b ... Compressor side casing 7 ... Expansion turbine 7a ... Turbine impeller 7b ... Turbine side casing 8 ... Second Heat exchanger 12 ... turbine unit housing 13 ... main shafts 13a, 13b ... thrust plate 14 ... spindle housings 15, 16 ... rolling bearing 15a ... inner ring 15b ... outer ring 17 ... electromagnet 18 ... sensor 19 ... magnetic bearing controller 20 ... rolling bearing Support portions 21A, 22A ... Sleeve (non-contact seal)
21B, 22B ... elastic support member 39 ... motor 39a ... motor rotor 39b ... motor stator A ... non-contact seal gap Aa ... small diameter side circumferential surface Ab ... gap large diameter side circumferential surface D1 ... inner diameter D2 ... outer diameter D3 ... outer diameter D4 ... inner diameter S ... bearing space

Claims (8)

The main shaft supporting the compressor side impeller and the turbine side impeller is supported by using both a rolling bearing and a magnetic bearing installed in the turbine unit housing, and a pressure difference between the compressor side impeller and the turbine side impeller with respect to the rolling bearing. A compression-expansion turbine system in which a non-contact seal is provided between a turbine unit housing and a main shaft adjacent to the high pressure side of the atmospheric pressure around the main shaft.
The radial positional relationship between the gap of the non-contact seal and the end of the bearing space formed between the inner and outer rings of the rolling bearing adjacent to the non-contact seal is a positional relationship in which no overlapping range occurs. A compression-expansion turbine system characterized.   2. The non-contact seal in the inner ring of the rolling bearing according to claim 1, wherein an inner diameter of the large-diameter side of the small-diameter and large-diameter peripheral surfaces facing each other constituting the clearance of the non-contact seal is A compression-expansion turbine system that is smaller than the outer diameter of adjacent end faces.   The outer diameter of the peripheral surface on the small diameter side of the peripheral surfaces on the small diameter side and the large diameter side that are opposed to each other and that constitutes the gap of the noncontact seal is the noncontact seal in the outer ring of the rolling bearing. A compression-expansion turbine system that is larger than the inner diameter of adjacent end faces.   4. The rolling bearing according to claim 1, wherein the rolling bearing supports a radial load, the magnetic bearing supports one or both of an axial load and a bearing preload, and the electromagnet of the magnetic bearing Mounted on the turbine unit housing so as to face the flange-shaped thrust plate made of a ferromagnetic material provided on the main shaft in a non-contact manner, and drives the compressor side impeller by the power generated by the turbine side impeller Is a compression-expansion turbine system.   5. The rigidity of the composite spring according to claim 4, further comprising a controller that controls the electromagnet in accordance with an output of a sensor that detects a force in the axial direction of the main shaft, and is formed by the rolling bearing and a support system for the rolling bearing. A compression / expansion turbine system having a relationship in which a value is greater than a negative stiffness value of the electromagnet.   4. The rolling bearing according to claim 1, wherein the rolling bearing supports a radial load, the magnetic bearing supports one or both of an axial load and a bearing preload, and the electromagnet of the magnetic bearing It is attached to the turbine unit housing so as to face the flange-shaped thrust plate made of a ferromagnetic material provided on the main shaft in a non-contact manner, and a motor rotor is provided on the thrust plate, and the motor stator is connected to the motor rotor. A compression-expansion turbine system that is disposed in the turbine unit housing so as to oppose and drives the compressor-side impeller by one or both of the power of the motor and the power generated by the turbine-side impeller.   7. The rigidity of the composite spring according to claim 6, further comprising a controller that controls the electromagnet according to an output of a sensor that detects a force in the axial direction of the main shaft, and is formed by the rolling bearing and a support system of the rolling bearing. A compression-expansion turbine system having a relationship in which a value is larger than a negative stiffness value of a synthetic spring formed by the electromagnet and the motor.   8. The compression / expansion turbine system according to claim 1, wherein the compression / expansion turbine system compresses the incoming air by a compressor of the turbine unit, cooling by another heat exchanger, and heat insulation of the turbine unit by the expansion turbine. Used in an air cycle refrigeration cooling system that sequentially performs expansion or compression by pre-compression means, cooling by a heat exchanger, compression by a compressor of a turbine unit, cooling by another heat exchanger, and adiabatic expansion of the turbine unit by an expansion turbine. Compressed and expanded turbine system.

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