JP2019019671A - Screw compressor - Google Patents

Abstract

To provide a screw compressor which can perform displacement control, ranging from a rated load to a minimum load, by use of a slide valve, can further reduce the minimum load at which a compression operation can be performed, and also achieves high efficiency at the rated load.SOLUTION: A screw compressor comprises: a casing 3, 4 that accommodates a screw rotor and has an intake port and an axial discharge port 16; a slide valve 13 that is axially disposed, in a freely reciprocating manner, at a compression-chamber-side meshing part of the screw rotor and changes the load rate by adjusting the timing for starting compression so as to control displacement; and a movable member 29, 30 that constitutes a portion of the axial discharge port and is configured such that the timing for discharging compressed gas can be changed by moving this portion of the axial discharge port.SELECTED DRAWING: Figure 5

Description

  The present invention relates to a screw compressor, and is particularly suitable for a constant speed type refrigerating screw compressor in which load control is performed by a slide valve (capacity control valve) that changes a suction flow rate.

  Constant-speed refrigeration (including refrigeration and air conditioning) screw compressors have high performance and high reliability, so they are used for relatively medium to large capacity chiller units in the food refrigeration market. It has been. In the chiller unit, capacity control with respect to the load is indispensable, and recently, an inverter screw compressor that changes the number of rotations of the compressor according to the load has been released.

  However, a constant speed screw compressor that performs load control using a slide valve is in demand because it is inexpensive. In a constant speed type screw compressor, a slide valve is provided in a main casing of the compressor, and the slide valve is provided at a meshing intersection portion on the compression chamber side of the screw rotor. This slide valve has a suction port formed on the suction side and a radial discharge port formed on the discharge side, and is configured to be slidable in the axial direction.

  A D casing (discharge casing) is attached to the discharge side of the main casing, and the D casing is provided so as to cover the discharge end face of the screw rotor. An axial discharge port is formed in the D casing.

The compression start timing for compressing the refrigerant in the screw rotor groove is determined by the suction port position of the slide valve. Further, the discharge start timing of the compressed refrigerant is determined by the discharge port that opens quickly among the radial discharge port provided in the slide valve and the axial discharge port provided in the D casing. Discharge starts from the discharge port on the opening side.
Examples of conventional screw compressors include those described in Patent Document 1 (Japanese Patent Laid-Open No. 2008-75618).

JP 2008-75618 A

In the above conventional one, the opening time of the axial discharge port provided in the D casing is the design volume ratio (at rated load) obtained from the design pressure ratio π _i (discharge pressure / suction pressure) at the rated load (load factor 100%). If the volume ratio of “rotor groove volume at the start of compression / rotor groove volume at the time of discharge”) V _i is set to 2.7, for example, the slide valve is in the discharge direction when the minimum load is, for example, a load factor of 25%. The radial discharge port that is slid and moved together with the slide is maintained at the opening position that maintains the design volume ratio V _i 2.7, but the axial discharge port provided in the D casing is compressed. Before starting, communication with the axial end portion (discharge side end portion) of the rotor groove (communication with the suction port side) is started, which causes a problem that the compression operation is not performed. .

Therefore, the conventional and the opening timing of the axial discharge ports provided in the D casing, wherein so as not to cause the suction port side and communicating, the design volume ratio V _i is set to be 4.0 or more ing. However, at the rated load (100% load factor), setting such that the opening timing of the design volume ratio V _i of the axial discharge ports are 4.0 or more, than the design capacity ratio V _i to those which 2.7 Therefore, there is a problem that the efficiency at the rated load decreases.

  It is an object of the present invention to control a capacity from a rated load to a minimum load with a slide valve, to make the minimum load capable of compression operation smaller, and to obtain a high efficiency at the rated load To get a compressor.

  In order to achieve the above object, the present invention accommodates a screw rotor, is provided in a casing having a suction port and an axial discharge port, and a meshing portion on the compression chamber side of the screw rotor so as to be capable of reciprocating in the axial direction. A slide valve that controls the capacity by changing the load factor by adjusting the compression start time and a part of the axial discharge port are formed, and a compressed gas is discharged by moving a part of the axial discharge port. A movable member configured to be able to change the time is provided.

  According to the present invention, it is possible to control the capacity from the rated load to the minimum load by the slide valve, to further reduce the minimum load that can be compressed, and to obtain high efficiency at the rated load. The effect which can obtain a compressor is acquired.

The horizontal sectional view which shows Example 1 of the screw compressor of this invention. The longitudinal cross-sectional view of the screw compressor shown in FIG. In an end view of a D casing with an axial discharge port design volume ratio V _i is 5.0, drawing also shows the rotor tooth located at the ejection timing relative to the axial discharge port. In an end view of a D casing with an axial discharge port design volume ratio V _i is 2.7, drawing also shows the rotor tooth located at the ejection timing relative to the axial discharge port. In an end view of a D casing with a movable axial discharge port in the first embodiment and shows a state where the volume ratio V _i moves the movable member to the position where the 5.0. In an end view of a D casing with a movable axial discharge port in the first embodiment and shows a state where the volume ratio V _i moves the movable member to the position where the 2.7. The longitudinal cross-sectional view of the female rotor side movable member vicinity in the female rotor part shown in FIG. 5, FIG. The load characteristic figure explaining the relationship between the load factor and efficiency in a screw compressor. The hydraulic system diagram for driving a slide valve and the movable member of a movable axial discharge port.

  Hereinafter, specific examples of the screw compressor of the present invention will be described with reference to the drawings. In each figure, the part which attached | subjected the same code | symbol has shown the part which is the same or it corresponds.

  The overall configuration of Embodiment 1 of the screw compressor of the present invention will be described with reference to FIGS. FIG. 1 is a horizontal sectional view of the screw compressor, and FIG. 2 is a longitudinal sectional view of the screw compressor shown in FIG. The screw compressor of the present embodiment is a constant speed type and for refrigeration.

  In these drawings, reference numeral 1 denotes a male rotor constituting a screw rotor of a screw compressor, 2 denotes a female rotor, and the male rotor 1 and the female rotor 2 are housed in a state where they are engaged with each other in a main casing 3. A D casing (discharge casing) 4 is connected to the discharge side of the main casing 3, and a motor casing 5 is connected to the suction side. The main casing 3, the D casing 4, and the motor casing 5 may be collectively referred to as a casing.

  A motor 6 for driving the male rotor 1 is installed in the motor casing 5. Further, the motor casing 5 is provided with a suction port 5a for sucking refrigerant gas of a refrigeration cycle (not shown) and a gas strainer 7 provided in the portion of the suction port 5a.

  8 is a suction port provided in the main casing 3, 9 is provided in the D casing, a male rotor discharge-side bearing for supporting the discharge-side shaft portion of the male rotor 1, and 10 is provided in the main casing 3. A male rotor suction side bearing that supports the suction side shaft portion of the male rotor 1, 11 is a female rotor discharge side bearing that is provided in the D casing and supports the discharge side shaft portion of the female rotor 2, and 12 is the main casing. 3 is a female rotor suction-side bearing that supports the suction-side shaft portion of the female rotor 2. The male rotor 1 and the female rotor 2 are rotatably supported by the bearings 9-12.

  The suction-side shaft portion of the male rotor 1 is directly connected to the motor 6, so that the male rotor 1 is rotated by the rotation of the motor 6, and the refrigerant in the rotor groove is compressed by the follower of the meshing female rotor. It is configured.

  Reference numeral 13 denotes a slide valve (capacity control valve) configured to perform load control by changing the suction flow rate. The slide valve 13 is a meshing intersection portion on the compression chamber side of the screw rotors 1 and 2 in the main casing 3. It is provided on the side where the male rotor 1 and the female rotor 2 are engaged with each other so as to be able to reciprocate. The slide valve 13 has a suction port (slide valve suction port) 14 formed on the suction side, and a radial discharge port (slide valve discharge port) 15 formed on the discharge side.

  The D casing 4 is provided so as to cover the discharge-side end face of the screw rotor, and an axial discharge port 16 is formed in the D casing 4 and discharged from the axial discharge port 16 and the radial discharge port 15. The compressed refrigerant gas is sent to the refrigeration cycle from the discharge port 4 a formed in the D casing 4.

  Reference numeral 17 denotes a slide valve rod having one end connected to the slide valve 13, and the other end of the rod 17 is connected to a slide valve piston 18. The piston 18 is configured to reciprocate in a slide valve cylinder 19 attached to the D casing 4. The slide valve spring 20 is provided in the slide valve cylinder 19 and applies a force in a direction to move the slide valve 13 to the discharge side to the slide valve piston 18.

  Next, the flow of the refrigerant will be described. The refrigerant gas of the refrigeration cycle sucked through the suction port 5a and the gas strainer 7 provided at the end of the motor housing 5 passes through the gap between the motor casing, the stator, the stator, and the rotor in the motor 6. It passes through and flows into the suction port (suction chamber) 8 in the main casing 3. Further, when the sucked refrigerant gas passes through the motor 6, the motor 6 is cooled by the refrigerant gas.

  The refrigerant gas flowing into the suction port 8 is filled into the rotor grooves of the male rotor 1 and the female rotor 2 facing the suction port 8 through a slide valve suction port 14 provided in the suction port 8 and the slide valve 13. Is done. At this time, the refrigerant gas is filled in the rotor groove together with the oil flowing into the suction port 8 after lubricating the bearings.

  The refrigerant gas filled in the rotor groove reduces the volume of the rotor groove formed by the meshing of the male and female rotors 1 and 2 with the rotation of the male rotor 1 and the female rotor 2. The gas is compressed.

  The compressed refrigerant gas is discharged together with the oil in the radial direction from the radial discharge port 15 of the slide valve 13 and in the axial direction from the axial discharge port 16 provided in the D casing 4.

  The shapes of the slide valve suction port 14 and the slide valve discharge port (radial discharge port) 15 provided in the slide valve 13 are constant, but the slide valve 13 slides in the axial direction. Depending on the position in the axial direction, the timing at which the rotor groove starts to be compressed and the timing at which the discharge is started vary. On the other hand, the conventional axial discharge port 16 provided in the D casing 4 is fixed at a certain position and shape, and therefore, according to the rotation of the screw rotor regardless of the position of the slide valve 13. It was configured to always open at a certain position (time).

Next, the configuration of the conventional axial discharge port 16 will be described with reference to FIGS. 3 and 4. Figure 3 is a end view of the D casing with an axial discharge port design volume ratio V _i is 5.0, drawing also shows the rotor tooth located at the ejection timing relative to the axial discharge ports, Figure 4 is designed in an end view of a D casing volume ratio V _i is provided with an axial discharge port to which 2.7 diagrams are also shown the rotor tooth located at the ejection timing relative to the axial discharge port. The shape and opening timing of the axial discharge port provided in the D casing 4 will be described with reference to these drawings.

FIG. 3 shows an axial discharge port where the design volume ratio V _i is 5.0. That is, the load factor of 100% indicates the shape of the axial discharge port volume ratio V _i becomes 5.0 when the (rated load), when operating at rated load, the volume ratio is 5.0 It is set to open when it becomes. Further, in FIG. 3, for reference, the tooth profile of the screw rotor when the axial discharge port is in the opening timing is also indicated by a one-dot chain line.
The axial discharge port 16 provided in the D casing 4 includes a male rotor side axial discharge port 16a and a female rotor side axial discharge port 16b.

  The basic male rotor-side axial discharge port 16a has a tooth tip circle (tooth tip diameter) 21 and a tooth root circle (tooth root diameter) 22 of the male rotor 1 (in the drawing, a slightly larger diameter is considered in consideration of leakage). , And a male rotor reverse surface tooth profile 23, and a shape in which an arc is connected between the tip circle 21 and the reverse surface tooth profile 23, and between the reverse surface tooth profile 23 and the root circle 22, respectively. It is configured.

  The female rotor axial discharge port 16b is formed of a tooth tip circle (tooth tip diameter) 24, a tooth root circle (tooth root diameter) 25, and an advancing surface tooth profile 26 of the female rotor 2, and the tooth tip circle 24 and the advance tip are formed. The surface tooth profile 26 and the advancing surface tooth profile 26 and the root circle 25 are also connected to each other by arcs. Further, the shape of the meshing portion of the axial discharge port 16 is adjacent between the tooth tip circles 21 ′ and 24 ′ of the male rotor 1 and the female rotor 2 and the tooth tip circles 21 ′ and 24 ′. The root circles 22 and 25 are also connected by arcs.

The opening timing of the axial discharge port 16 is determined by the positions in the circumferential direction of the reverse surface tooth profile 23 of the male rotor 1 and the forward surface tooth profile 26 of the female rotor 2. That is, the male rotor 1, the female rotor 2, the relative groove volume of the main casing 3 and when the suction in the grooves volume formed by the D casing 4 (the beginning confinement), the volume ratio V _i is 5.0 The tooth profile position of each of the rotors 1 and 2 when the groove volume is the opening position of the axial discharge port 16 is used.
Therefore, when the opening timing of the axial discharge port 16 changes, the circumferential position of the reverse face tooth profile 23 of the male rotor 1 and the forward face tooth profile 26 of the female rotor 2 also changes.

FIG. 4 shows an axial discharge port with a design volume ratio V _i of 2.7. That is, the load factor of 100% indicates the volume ratio V _i is the shape of the axial discharge ports to be 2.7 when the (rated load), when operating at rated load, the volume ratio is 2.7 It is set to open when it becomes. For reference, the tooth profile of the screw rotor when the axial discharge port is at the opening time is also indicated by a one-dot chain line in FIG.

In the axial discharge port where the design volume ratio V _i shown in FIG. 4 is 2.7, the male rotor-side axial discharge port 16a and the female rotor-side axial discharge port 16b are compared with the case of FIG. The female rotor advance surface tooth profile 26 is in a position that opens earlier than the rotation of the rotor.

In the case of the rated load (100% load factor) it is found the design volume ratio V _i shown in FIG. 4 to set the axial discharge port at a position of 2.7, preferable from the viewpoint of performance. However, when the axial discharge port is set at a position where the design volume ratio V _i is 2.7, when the slide valve 13 is slid to the discharge side at this position, for example, when the load factor is 25%, the screw rotor groove ( Since the compression start time for compressing the refrigerant in the rotor groove) is delayed, the axial end portion (discharge side end portion) of the rotor groove before the start of compression communicates with the axial discharge port 16. That is, since the suction port 8 side and the axial discharge port 16 communicate with each other through the rotor groove, compression cannot be performed.

Therefore, in the conventional screw compressor, because even move the slide valve 13 to a load 25% of the positions to allow compression, as shown in FIG. 3, the design volume ratio V _i becomes 5.0 Thus, the axial discharge port 16 was formed. However, when the opening timing of the axial discharge port is set so that the design volume ratio V _i is 5.0, the rated load (load factor 100%) is higher than that when the design volume ratio V _i is 2.7. The efficiency in time decreases.

  In this embodiment, in order to solve this problem, as shown in FIGS. 5 and 6, the portions of the male rotor backward surface tooth profile 23 and the female rotor forward surface tooth profile 26 that determine the opening timing of the axial discharge port 16 are respectively set. It is composed of a movable member. That is, a plate-shaped male rotor side movable member 29 having a male rotor reverse surface tooth profile 23 at the end (tip portion) and a plate-shaped female rotor side movable member having a female rotor advance surface tooth profile 26 at the end (tip portion). The member 30 is provided on the discharge side end surface portion of the D casing 4.

  That is, in this embodiment, the male rotor reverse surface tooth profile 23 and the female rotor advance surface tooth profile 26 forming the axial discharge port 16 are arranged on the plate-shaped male rotor side having the male rotor reverse surface tooth profile 23 at the end. The movable axial discharge port is formed by a movable member 29 and a plate-shaped female rotor side movable member 30 having a female rotor advance surface tooth profile 26 at its end.

  The male rotor-side movable member 29 and the female rotor-side movable member 30 are respectively provided with a male-rotor-side movable member guide groove 27 and a female-rotor-side movable member guide groove 28 provided on the discharge-side end surface portion of the D casing 4. And is configured to be slidable in the tangential direction of the rotor along the movable member guide grooves 27 and 28 that are linearly formed.

In the present embodiment, as described above, the axial discharge port 16 is configured as a movable axial discharge port formed by using the movable members 29 and 30, so that when operating at a rated load (load factor: 100%), as shown in FIG. 6, as designed volume ratio V _i is the axial discharge port 16 to be 2.7 (the value determined from the design pressure ratio) or less, to move the movable member 29, 30.

Further, during operation at the minimum load (for example, a load factor of 25%), as shown in FIG. 5, the design volume ratio V _i is, for example, 5.0 (the design volume ratio V _{L that} is larger than the design volume ratio V _i). The movable members 29 and 30 are moved so that the axial discharge port 16 becomes. Thereby, it is possible to prevent the axial discharge port from communicating with the suction port through the rotor tooth groove even during operation at the minimum load.

Incidentally, when the above minimum load at a load of less than 100%, the design volume ratio V _i is, for example, so that the appropriate value in the range of 2.7 to 5.0, the load factor (the position of the slide valve 13) Accordingly, the movable members 29 and 30 may be moved to appropriate positions. That is, the movable member may be moved so that the design volume ratio increases as the load factor decreases in accordance with the load factor controlled by the slide valve.

Further, the slide valve 13 has the radial discharge port 15 is provided (see FIG. 2) is provided at a position which opens when the volume ratio V _i becomes 2.7, the radial discharge port 15 regardless of the load factor (the slide valve position), always open when the volume ratio V _i becomes 2.7. Accordingly, it is preferable to control the movement of the movable members 29 and 30 so that the axial discharge port 16 is also opened as much as possible in accordance with the opening timing of the radial discharge port 15.

  As described above, by moving the movable members 29 and 30 according to the load factor (slide valve position), even at the minimum load (for example, a load factor of 25%), the axial end portion ( It is possible to prevent the discharge side end portion) from communicating with the axial discharge port 16. That is, by moving the movable member so that the axial discharge port has a design volume ratio that does not communicate with the suction port through the rotor tooth groove, it is possible to perform compression even at the minimum load.

At the rated load, the movable members 29 and 30 are moved so that the axial discharge port 16 opens with a design volume ratio V _i of 2.7, so that the design volume ratio V _i at the rated load is 5 Compared to the case of using an axial discharge port of 0.0, the effect of improving the efficiency at the rated load can be obtained.

  The male rotor side movable member 29 and the female rotor side movable member 30 are each connected to a movable member piston 32 accommodated in a movable member cylinder 33 via a movable member rod 31. A spring 34 is provided in the movable member cylinder 33 to urge the piston 32 in a direction that advances the opening timing of the axial discharge port 16, and further, the opening timing of the axial discharge port 16 is delayed. In order to operate the piston 32, a hydraulic inlet path 38 for supplying hydraulic pressure into the cylinder 33 is provided.

  Also, hydraulic discharge paths 35 and 36 for discharging the hydraulic pressure in the cylinder 33 are provided, and the reaction force of the spring 34 is discharged by discharging the hydraulic pressure in the cylinder 33 from the hydraulic pressure discharge path 35 or 36. Thus, the movable members 29 and 30 can be controlled so that the opening timing of the axial discharge port 16 is advanced. When the hydraulic pressure is discharged from the hydraulic pressure discharge passage 36, the opening timing of the axial discharge port 16 can be made earlier than when the hydraulic pressure is discharged from the hydraulic pressure discharge passage 35. Reference numeral 37 denotes a fixed housing for fixing the movable member cylinder 33.

  FIG. 7 is a longitudinal sectional view in the vicinity of the female rotor side movable member in the female rotor portion shown in FIGS. 5 and 6, and the seal structure in the portion of the female rotor side movable member 30 will be described with reference to FIG. .

  In FIG. 7, 28 is a movable member guide groove provided on the female casing 2 side of the D casing 4, and 30 is a female rotor side movable member that is inserted into the movable member guide groove 28 and is slidably accommodated. Since there is a gap between the movable member guide groove 28 and the female rotor side movable member 30, in order to prevent the refrigerant gas compressed from the discharge port side to the suction port side from leaking through this gap, In this embodiment, seal grooves 39 are provided on the upper and lower surfaces of the movable member guide groove 28, and seal members 40 are provided in the seal grooves 39.

The seal groove 39 is provided along the longitudinal direction of the female rotor movable member 30, and the cross-sectional shape thereof may be rectangular as shown in FIG. The cross-sectional shape of the sealing material provided in 39 may be circular as shown in FIG. 7, or a sealing material having an arbitrary cross-sectional shape such as a rectangle may be used.
Further, since the male rotor side movable member 29 side described above has the same configuration as the female rotor side movable member 30 described above, the description of the male rotor side movable member 29 side is omitted.

FIG. 8 is a load characteristic diagram illustrating the relationship between the load factor and the efficiency in the screw compressor. The effect of the present embodiment will be described with reference to FIG.
A curve a indicated by a solid line indicates a characteristic curve of a screw compressor using an axial discharge port where the design volume ratio V _i shown in FIG. 3 is 5.0. The efficiency when operating such a screw compressor at a load factor of 25% was point A, and the efficiency when operating at a rated load (load factor of 100%) was point B.

A curve b shown by a broken line shows a characteristic curve of a screw compressor using an axial discharge port where the design volume ratio V _i shown in FIG. 4 is 2.7. The efficiency when such a screw compressor is operated at the rated load (load factor 100%) is point C, which is improved compared to the case of curve a (point B). However, the efficiency when operating at a load factor of 25% is greatly reduced to the D point. This is because, as described above, communication between the axial end portion of the rotor groove before the start of compression and the axial discharge port 16 occurs.

A curve c indicated by a dotted line indicates a characteristic curve of the screw compressor of the present embodiment provided with the above-described movable axial discharge port. In the present embodiment, during operation at a rated load (100% load), the said movable member 29, 30 is moved, and controls so designed volume ratio V _i is the axial discharge port to which 2.7 It is possible to operate at the efficiency of the point C, and it is possible to perform a highly efficient operation with improved efficiency compared to the case of the curve a (point B). In the case of operation at a load factor of 25% is said movable member 29, 30 is moved, by controlling so designed volume ratio V _i is the axial discharge port of 5.0, the efficiency point It becomes possible to improve to E.

The dotted line c to the present exemplary embodiment, when the minimum load or driving the load factor less than 100%, the design volume ratio V _i at the rated load, the appropriate value in the range of 2.7 to 5.0 Thus, it is an example at the time of moving the said movable members 29 and 30 to an appropriate position according to a load factor (slide valve position). That is, it is a characteristic curve showing the efficiency when the movement of the movable members 29 and 30 is controlled so that the axial discharge port 16 opens almost simultaneously with the opening timing of the radial discharge port 15.

  In general, a screw compressor provided with a slide valve is often controlled in steps of, for example, 100%, 75%, 50%, and 25%. Therefore, according to the movement of the slide valve 13, By moving the movable members 29 and 30, the axial discharge port 16 can be configured to open almost simultaneously with the opening timing of the radial discharge port 15.

Further, the vicinity of the curve a and b intersect, i.e. the load rate is up to 75% to 85%, the design volume ratio V _i is from 2.6 to 2.9 and comprising an axial discharge port and so as to the movable member 29, 30 controls, in the range of the load factor smaller load factor than control the movable member 29, 30 so that the axial discharge port design volume ratio V _i becomes 4.0 to 5.0 in two steps You may make it do. When controlling in this way, the control becomes simple, and also in efficiency, the efficiency along the curve b is obtained on the high load side (for example, load factor 75 to 100%), and on the low load side (for example, 75% or less). Efficiency along the curve a can be obtained.

FIG. 9 shows a hydraulic system diagram for driving the slide valve and the movable member of the movable axial discharge port.
In FIG. 9, 13 is the above-described slide valve, 29 and 30 are movable members constituting the above-described movable axial discharge port, 29 is a male rotor-side movable member, and 30 is a female rotor-side movable member. In the present embodiment, the slide valve 13 and the two movable members 29 and 30 can be driven simultaneously by using a single hydraulic system.

  That is, the slide valve 13 and the movable members 29 and 30 are each driven by a hydraulic mechanism including a cylinder, a piston that reciprocates in the cylinder, and a spring provided in the cylinder. A hydraulic inflow passage having an electromagnetic valve and a plurality of hydraulic discharge passages having an electromagnetic valve are connected to each other. The piston is moved in the direction of compressing the spring by supplying hydraulic pressure into each cylinder, and the piston is slid in the reverse direction by the spring force of the spring by releasing the hydraulic pressure in the cylinder. It is configured. Therefore, the slide valve and the movable member can be positioned at a plurality of predetermined positions by controlling the opening and closing of the electromagnetic valves. Hereinafter, this configuration will be described in detail with reference to FIG.

  41 is a pressure oil container to which the gas pressure of the refrigerant gas on the discharge side of the screw compressor is applied. The oil in the pressure oil container 41 is passed through an oil strainer 42 provided in the pressure oil container 41. The pressure is supplied to the hydraulic pressure supply pipe 43 by the pressure difference from the pressure on the suction port 8 (see FIG. 2) side. The hydraulic pressure supply pipe 43 is branched into three systems of hydraulic inflow passages 38 via electromagnetic valves 44 and 45, and by opening and closing the electromagnetic valves 44 and 45, the slide valve cylinder 19 and the male rotor side movable member 29 can be moved. The hydraulic pressure is applied to the member cylinder 33 and the movable member cylinder 33 of the female rotor side movable member 30, or the hydraulic pressure is released.

That is, by opening the electromagnetic valve 44, the slide valve 13 and the slide valve piston 18 and the slide valve are arranged in a direction in which the slide valve spring 20 provided in the cylinder 19 is compressed (in the direction of increasing the capacity). It is moved through the rod 17. Further, by opening the electromagnetic valve 45, the male rotor-side movable member 29 and the female rotor-side movable member 30 are movable in the direction in which the movable member spring 34 is compressed (the direction in which the design volume ratio V _i increases). It is moved via the member piston 32 and the movable member rod 31.

  The slide valve cylinder 19 is provided with two hydraulic discharge passages 50 and 51, and the hydraulic discharge passages 50 and 51 are provided with electromagnetic valves 46 and 47, respectively. By opening the electromagnetic valve 46, the oil in the cylinder 19 is discharged to the suction port 8 side, and the piston 18 moves to the position of the hydraulic pressure discharge path 50 by the spring force of the spring 20. The position of the slide valve 13 at this time is, for example, a position where the load factor is 50%. Further, by opening the electromagnetic valve 47, the piston 18 moves to the position of the hydraulic pressure discharge path 51 by the spring force of the spring 20. The position of the slide valve 13 at this time is, for example, a position where the load factor is 25%. Further, when both the solenoid valves 46 and 47 are closed, the piston 18 is moved in the direction that compresses the spring 20 most, and the load factor becomes 100%.

Two hydraulic discharge passages 35 and 36 are also provided in each cylinder 33 of the movable members 29 and 30, and electromagnetic valves 48 and 49 are also provided in these hydraulic discharge passages 35 and 36, respectively. By opening the electromagnetic valve 48, the oil in the cylinder 33 is discharged to the suction port 8 side, and the piston 32 moves to the position of the hydraulic pressure discharge path 35 by the spring force of the spring 34. Position of the movable member 29 at this time, intermediate load (for example load factor 50%) suitable design volume ratio _{V i} (e.g. _{V i} corresponding to 4.0, such as between 2.7 to 5.0 Value). Further, by opening the electromagnetic valve 49, the piston 32 moves to the position of the hydraulic pressure discharge path 36 by the spring force of the spring 34. The positions of the movable members 29 and 30 at this time are positions where an appropriate design volume ratio V _i (for example, V _i is 2.7) corresponding to the rated load (load factor 100%). Furthermore, when both the solenoid valves 48 and 49 are closed, the piston 32 is moved in the direction in which the spring 34 is most compressed, and an appropriate design volume ratio V _i corresponding to the minimum load (for example, a load factor of 25%). (For example, V _i 5.0).

Therefore, when operating at a load factor of 25%, the solenoid valve 44 is closed and the solenoid valve 47 is opened, so that the hydraulic pressure of the slide valve cylinder 19 is released, and the slide force is applied by the spring force of the slide valve spring 20. Slide the valve 13 to the discharge side. Further, by opening the solenoid valve 45 and closing the solenoid valves 48 and 49, hydraulic pressure is applied to the movable member cylinder 33, and the opening timing of the axial discharge port is delayed (the design volume ratio V _i is large). The movable members 29 and 30 can be slid in the same direction) to form an axial discharge port having a design volume ratio (for example, V _i 5.0) corresponding to a load factor of 25%.

When operating at a load factor of 100%, conversely, by opening the solenoid valve 44 and closing the solenoid valves 46 and 47, hydraulic pressure is supplied to the slide valve cylinder 19 and the slide valve spring 20 is compressed. The slide valve 13 is slid to the suction side. Further, by closing the electromagnetic valve 45 and opening the electromagnetic valve 49, the hydraulic pressure in the movable member cylinder 33 is released, and the movable members 29 and 30 advance the opening timing (the design volume ratio V _i is The axial discharge port with a design volume ratio (for example, V _i 2.7) corresponding to a load factor of 100% can be obtained.

  In the example shown in FIG. 9, the example in which the position of the slide valve 13 and the positions of the movable members 29 and 30 are controlled in three steps (load factor 100%, 50%, 25%) has been described. Alternatively, it is possible to control in four steps or more.

For example, when the control is performed in four stages, the position of the slide valve 13 is set to, for example, 100%, 75%, 50%, 25%, and the positions of the movable members 29, 30 are also corresponding to the four stages. The design volume ratio V _i of the movable axial discharge port is, for example, 2.7, 3.7, 4.0 (or 5), respectively, corresponding to the load factors of 100%, 75%, 50%, and 25%. .0), 5.0.

  According to the embodiment of the present invention described above, in the case where the capacity is controlled from the rated load to the minimum load by the slide valve 13, it is possible to further reduce the minimum load that can be compressed, and at the time of the rated load. There is an effect that a screw compressor that can be operated with high efficiency can be obtained.

For example, when operating at a load ratio of 25 percent, said movable member 29, 30 is moved, the design volume ratio V _i is by an axial discharge port to be 4.0 to 5.0, the load factor 25 % Or less, it is possible to operate at a low load. Furthermore, when operating at rated load (100% load factor) is said movable member 29, 30 is moved, by the design volume ratio V _i is the axial discharge port to be 2.7, with high efficiency An operable screw compressor can be obtained.

In addition, this invention is not limited to the Example mentioned above, Various modifications are included. For example, in the above-described embodiment, the case of a refrigeration screw compressor that compresses refrigerant gas has been described. However, the present invention is not limited to a refrigeration screw compressor, but a screw that compresses other gas such as air. The same applies to a compressor.
The above-described embodiments have been described in detail for easy understanding of the present invention, and are not necessarily limited to those having all the configurations described.

1: male rotor, 2: female rotor,
3 to 5: casing (3: main casing, 4: D casing, 5: motor casing), 4a: discharge port, 5a: suction port,
6: motor, 7: gas strainer, 8: suction port,
9: Male rotor discharge side bearing, 10: Male rotor suction side bearing,
11: Female rotor discharge side bearing, 12: Female rotor suction side bearing,
13: Slide valve,
14: Suction port (slide valve suction port),
15: Discharge port (slide valve discharge port),
16: Axial discharge port (movable axial suction port),
16a: Male rotor side axial discharge port, 16b: Female rotor side axial discharge port,
17: Slide valve rod, 18: Slide valve piston,
19: Slide valve cylinder, 20: Slide valve spring,
21, 21 ': male rotor tip circle (tooth diameter), 22: male rotor root circle,
23: Male rotor reverse surface tooth profile,
24, 24 ': female rotor tip circle (tooth tip diameter), 25: female rotor root circle,
26: female rotor forward face tooth profile,
27, 28: movable member guide groove,
29: Male rotor side movable member, 30: Female rotor side movable member,
31: Movable member rod, 32: Movable member piston, 33: Movable member cylinder,
34: movable member spring, 35, 36: hydraulic pressure discharge path,
37: Fixed housing, 38: Hydraulic inflow path,
39: Seal groove, 40: Seal material,
41: Pressure oil container, 42: Oil strainer,
43: Hydraulic supply pipe, 44 to 49: Solenoid valve, 50, 51: Hydraulic discharge path.

Claims (9)

A casing having a screw rotor and a suction port and an axial discharge port;
A slide valve that is reciprocally movable in the axial direction at the meshing portion on the compression chamber side of the screw rotor, and controls the capacity by changing the load factor by adjusting the compression start timing;
A screw comprising: a movable member configured to change a timing at which compressed gas is discharged by forming a part of the axial discharge port and moving a part of the axial discharge port Compressor.   2. The screw compressor according to claim 1, wherein the slide valve is moved so as to control the capacity between a rated load and a minimum load, and becomes less than a design volume ratio obtained from a design pressure ratio during operation at the rated load. The movable member is moved so as to be an axial discharge port, and the movable member is moved so that the axial discharge port has a design volume ratio that does not communicate with the suction port through the rotor tooth groove during operation at the minimum load. A screw compressor characterized by   3. The screw compressor according to claim 2, wherein the movable member is moved according to a load factor controlled by the slide valve so that a design volume ratio increases as the load factor decreases. Compressor.   2. The screw compressor according to claim 1, wherein the slide valve is provided with a radial discharge port, and the movable member is configured to open the axial discharge port almost simultaneously with the opening timing of the radial discharge port. A screw compressor characterized by controlling the movement of the screw.   2. The screw compressor according to claim 1, wherein when the load factor is 75 to 85%, the movable member is controlled to be an axial discharge port having a design volume ratio of 2.6 to 2.9, and the load A screw compressor, wherein the movable member is controlled in two stages so that an axial discharge port having a design volume ratio of 4.0 to 5.0 is obtained in a range of a load factor smaller than the rate.   2. The screw compressor according to claim 1, wherein the movable member is a male rotor-side movable member in which a tip portion forming the axial discharge port is configured in a shape along a reverse surface tooth profile of the male rotor; The distal end portion forming the axial discharge port is constituted by a female rotor side movable member having a shape along the advance surface tooth profile of the female rotor, and both the movable members are moved in the tangential direction of the rotor shaft. A screw compressor characterized by comprising. The screw compressor according to claim 6, wherein the casing includes a main casing that houses the screw rotor, and a D casing that is attached to the main casing so as to cover a discharge-side end surface of the screw rotor,
The male rotor side movable member and the female rotor side movable member are accommodated in a male rotor side movable member guide groove and a female rotor side movable member guide groove provided on the discharge side end surface portion of the D casing. A screw compressor configured to be slidable in a tangential direction of a rotor shaft along a groove.   The screw compressor according to claim 7, wherein a seal groove is provided on each of the upper and lower surfaces of the movable member guide groove along the longitudinal direction of the movable member, and a seal material is provided in each seal groove, A screw compressor, wherein a gap between the movable member and the movable member guide groove is sealed. The screw compressor according to claim 1, wherein the slide valve and the movable member are configured to be controlled in conjunction with each other using a single hydraulic system,
The slide valve and the movable member are each driven by a hydraulic mechanism including a cylinder, a piston that reciprocates in the cylinder, and a spring provided in the cylinder,
Each cylinder is connected to a hydraulic inlet passage having a solenoid valve and a plurality of hydraulic outlet passages having a solenoid valve, respectively.
The piston is moved in a direction to compress the spring by supplying hydraulic pressure into the cylinder, and the piston is slid in the reverse direction by the spring force of the spring by releasing the hydraulic pressure in the cylinder. The screw compressor is configured to position the slide valve and the movable member at a plurality of predetermined positions by controlling opening and closing of the electromagnetic valves.

Images