JP4294212B2 - High pressure screw compressor - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a screw compression apparatus used for high-pressure compression of a gas fluid, and particularly in a screw compression apparatus that is operated at a very high discharge pressure when performing a supercritical operation of a CO ₂ refrigerant, The present invention relates to a screw compression device in which an excessive thrust force is not applied to the rotor shaft support portion in the reverse direction.
[0002]
[Prior art]
As shown in FIG. 5, a conventional screw compressor has a male rotor 5 and a female rotor 6 arranged in parallel in a casing comprising a rotor casing 1 and bearing housings 2 and 3 provided on both sides thereof. Each of the rotor shafts receives a radial load via bearings 7 and 8, and receives a thrust load by thrust bearings 9 and 9 ′. The shaft end of the discharge-side shaft of the male rotor 5 is connected to an electric motor or the like. To be driven. The shaft end side chamber 34 is closed by a cover 37 is formed, the shaft end portion of the suction side shaft of the male rotor 5 in the indoor 34 balance piston 30 is mounted, et al are also the discharge side of the thrust bearing A mechanical shaft seal 10 ′ is provided on the outside.
[0003]
During operation of the screw compressor, the gas fluid is sucked from the suction port 17, compressed in the male and female rotors 5 and 6 that are engaged with each other, and discharged from the discharge port 18.
The gas fluid discharged from the discharge port 18 is guided to an oil separator indicated by reference numeral 21 in FIG. 3 or 4 showing an embodiment of the present invention, where each part mixed in the discharged gas fluid. The lubricating oil that has been lubricated is separated and the gas fluid is supplied to a required external load, and the separated lubricating oil is supplied to the oil supply port 15.
[0004]
The lubricating oil supplied to the oil supply port 15 is subjected to the pressure in the oil separator, that is, the discharge pressure of the gas fluid, and the chamber 34 applies the pressure of the balance piston 30 through the oil passages 32 and 33. Is further sent to the rotor chamber through the chamber 35 behind the shaft end of the female rotor 6, and the rotor is lubricated and discharged together with the discharged gas fluid, and is separated from the gas fluid by the oil separator 21. The outer periphery of the balance piston 30 is fitted to the inner periphery of a balance piston sleeve 31 fixed to the bearing housing 3 so as to freely move. The male rotor 5 is pushed to the discharge side, that is, leftward by the pressure in the chamber 34. The left thrust force reduces the thrust force that pushes the male rotor in the right direction due to the pressure difference between the discharge side and the suction side.
[0005]
In the above configuration, when the operation of the screw compressor is stopped, the compressed gas fluid on the discharge side leaks to the suction side from the gap between the rotor casing and the rotor, and the pressure difference between the discharge side and the suction side decreases rapidly. The chambers 34 and 35 of the rotor casing 1, the chambers 9 and 9 'in which the thrust bearings are mounted, and the chamber of the radial bearing 7 are equalized to a pressure sufficiently lower than the discharge pressure but lower than the discharge pressure. That is, as shown in FIG. 3 or 4, check valves 22 and 23 are provided at the gas fluid outlet of the oil separator 21 and the inlet 17 of the screw compressor, so that the oil separator The space from 21 to the suction port 17 of the screw compressor is a closed space, and when the operation is stopped, the pressure in the space is equalized by the leakage of the gas fluid from the discharge side to the suction side, and the discharge pressure and suction during operation Although it is slightly lower than the discharge pressure between the pressures, the pressure is equalized to a pressure sufficiently higher than the suction pressure.
[0006]
On the other hand, although the pressure in the oil separator 21 that is equalized slightly lower than the discharge pressure during the operation is applied to the balance piston chamber 34, the rotor chamber that is the destination of the lubricating oil in the seal chamber 34 is also provided. Since the pressure is equalized, the lubricating oil is not sent to the rotor chamber, and the pressure in the seal chamber 34 is also maintained at the equalized pressure.
Therefore, after the operation is stopped, the pressure on the discharge side and the suction side of the rotor is quickly equalized so that there is no pressure difference, and the thrust force in the suction side due to the gas fluid rapidly becomes zero, but the mechanical seal portion 10 'on the discharge side Is maintained at a pressure slightly lower than the discharge pressure during operation, the thrust bearing 9 has a pressure slightly lower than the discharge pressure in the chamber of the mechanical seal portion 10 ′, and the differential pressure of the atmospheric pressure causes a difference from that during operation. In the reverse direction, a thrust force in the discharge side direction is applied.
[0007]
[Problems to be solved by the invention]
In such a configuration, when the screw compressor is used as an ordinary compressor for a refrigerator, for example, in the case of a chiller using an R22 refrigerant, the suction pressure is about 4 kg / cm ² and the discharge pressure is about 15 kg / cm ² , Thrust force can be reduced by the balance piston.
However, the suction pressure by the operating conditions when performing supercritical operating CO ₂ as refrigerant 20 to 50 kg / cm ^2, discharge pressure becomes very high pressure and 100 kg / cm ² before and after.
When performing such a high pressure operation, in the conventional configuration as described above, when the operation is stopped, the suction by the gas fluid applied to the male rotor 5 due to the differential pressure between the discharge pressure and the suction pressure of the male rotor 5 is performed. Since the thrust force in the lateral direction (rightward direction) suddenly decreases to zero, while the pressure in the mechanical seal portion 10 ′ chamber is slightly lower than 100 kg / cm ² but is maintained at a pressure close to that, the male seal portion 10 ′ The thrust bearing 9 of the rotor 5 receives the force in the discharge side direction (left direction) due to the high pressure slightly lower than about 100 kg / cm ^{2 in the} mechanical seal portion 10 ′ chamber as it is in the thrust bearing 9 in the direction opposite to that during operation. It takes.
[0008]
The gap between the rotor-side end surface and the surface facing the rotor of the discharge-side bearing housing 2 is very small for improving the volumetric efficiency, so that if the thrust force in the discharge-side direction when the operation is stopped is excessive, Due to the deformation and damage of the thrust bearing portion, the both surfaces come into contact with each other and burnout is caused.
Since the discharge pressure is not so high in the conventional refrigerator, the thrust thrust in the discharge side when the operation is stopped can be adequately handled by the thrust bearing, but the discharge pressure is higher than in the conventional refrigerator such as a refrigerator using CO ₂ as a refrigerant. In refrigeration cycle applications where the pressure is much higher, the thrust thrust force in the discharge side becomes excessive when the operation is stopped, the thrust bearing is damaged, and the rotor end face is burned out.
[0009]
In the present invention, in view of the above problems, when a screw compressor used under a high discharge pressure is stopped, an excessively large thrust force is applied to the thrust bearing in a direction opposite to that during operation. Therefore, it is an object of the present invention to provide a screw compression apparatus that does not cause damage to the thrust bearing and burnout of the rotor side surface.
[0010]
[Means for Solving the Problems]
According to the first aspect of the present invention, there is provided a rotor casing for housing a rotor chamber in which a male rotor and a female rotor, which have shaft portions on both sides and mesh with each other, are rotatably supported in parallel, and suction ports on both ends of the rotor of the casing. And a screw compressor body comprising a screw compressor body having a discharge port, and an oil separator for separating lubricating oil mixed in the compressed fluid discharged from the discharge port.
The male rotor serving as a drive shaft is provided with a mechanical seal portion including a seal rotating member fixed to the shaft and a seal fixing member fixed to the casing on the discharge side and the suction side of the male rotor shaft. While opening both ends of the shaft to the outside air,
Forming a suction-side seal chamber for applying a supply hydraulic pressure from the oil separator on the surface opposite to the outside air side of the suction-side seal rotating member;
A discharge-side seal that sets the discharge-side seal rotating member to have a larger diameter than the suction-side seal rotating member and applies the supply hydraulic pressure from the oil separator to the surface opposite to the outside air side of the discharge-side seal rotating member Forming a chamber,
Reducing the thrust force in the discharge side direction due to the pressure difference between the pressure in the rotor chamber and the atmospheric pressure generated when the operation is stopped with the outside of the seal rotating members of the suction side seal chamber and the discharge side seal chamber open to the atmosphere ,
It is characterized in that the thrust force in the suction side direction due to the differential pressure between the discharge pressure and the suction pressure generated during operation is relieved by the hydraulic pressure in the both seal chambers.
In this case, as described in claim 5, it is preferable that the mechanical seal portion is provided on the outer side of the bearing portion that supports the shaft portions on the discharge side and the suction side of the male rotor shaft, respectively. It is preferable that a supply oil pressure adjusting means is provided in the middle of the passage where the supply oil pressure separated by the oil separator is applied to the seal chamber as described in 2.
[0011]
The adjusting means for adjusting the hydraulic pressure is preferably a throttle portion as described in claim 3 .
[0012]
Also, as in the 請 Motomeko 4, wherein the throttle valve is provided from the discharge side seal chamber to a supply path for supplying lubricating oil to the rotor chamber between the casing and the rotor, the pressure of the discharge side sealing chamber at the narrowed valve The thrust force applied in the suction side direction during the operation of the screw compressor is adjusted so that it can be relaxed .
[0013]
From FIG. 1, the formula of the thrust force based on the pressure is as follows.
P = -A2, P0 + (A2-A6), P2- (A4-A6), P2 + F- (A7-A4), Ps + (A7-A3), Pd + (A3-A5), P1- (A1-A5), P1 + A1 ・ P0
Here, since P1 and P2 are equalized during operation and during stoppage, P1 = P2, and Ps (suction pressure) is substantially the same pressure. For simplicity, P1 = P2 = Ps
P = (A1-A2) .P0 + (A2-A1) .P1 + F + (A7-A3). (Pd-P1) (A)
It becomes. Here, P0 is atmospheric pressure, and can be neglected sufficiently smaller than P1 in high-pressure applications such as carbon dioxide. Moreover, the area difference of the underlined part of Formula (A) is always A7−A3 ≧ 0 from the shape of the rotor, and does not become a direction to cancel the thrust force. The diameter of A7 is determined by the shape of the rotor, and the area difference is made as small as possible. However, as with the gas load F, it is determined by the pressure difference between the suction and discharge during operation. The formula is
P = (A2-A1) · P1 + F
It becomes. In addition, since the suction and discharge pressures are equalized even at the time of stop, Pd−Ps = 0, and naturally, from F = 0,
P = (A2-A1) · P1
It becomes. Therefore, the thrust force is determined by the areas A1 and A2 and the in-machine pressure P1 formed by the thrust force F applied in the suction side direction and the seal diameter due to the difference between the suction pressure and the discharge pressure of the elastic fluid during operation.
[0014]
The present invention reduces the load on the thrust bearing in the discharge direction due to the pressure difference between the compressor internal pressure and the atmospheric pressure at the time of stopping (eventually prevents the thrust bearing and rotor side surface from being burned out) and gas during operation. The purpose is to reduce the load on the thrust bearing by offsetting the thrust force applied in the suction direction due to the difference between the suction pressure and the discharge pressure of the fluid by using a seal attached to the suction side and the discharge side.
Regarding the former, a seal for releasing the atmosphere is attached to the seal on the drive shaft side (discharge side) and the suction side which is the opposite side, and the latter is an area formed by the seal diameter of both seals (FIG. 1). A1 and A2) are changed to reduce the thrust force applied in the suction side direction due to the difference between the suction pressure and the discharge pressure generated during operation.
[0015]
Therefore, in general, the in-machine pressures P1 and P2 in FIG. 1 in the later-described embodiment are the same pressure (equal pressure), so that they are formed by the lip inner diameters D1 and D2 regardless of the area formed by the diameters D3 and D4. Thrust load is generated depending on the ring area.
Therefore, seal chamber pressures P1 and P2 (normally both seal chamber pressures are equalized, P1 = P2), seal lip inner diameter areas A1 and A2, and the suction side direction due to the difference between the suction pressure and discharge pressure of the elastic fluid during operation Assuming that the thrust force F is applied, the load P applied to the thrust bearing is as follows.
During stop: P = (A2-A1) × P1 (1) equation During operation: P = F + (A2-A1) × P1 (2) equation
Here, when the load direction is + when the suction direction is +, F is always +. If A2 = A1, the load applied to the thrust bearing at the time of stop is P = 0, and only F acts on the thrust bearing during operation. Here, in order to reduce the load P applied to the thrust bearing during operation, the inner shape area (oil supply pressure area) of the seal lip is set to A1> A2. However, if A1> A2, the load applied to the thrust bearing at the time of stopping increases, so the set value (A2-A1) must be determined to such an extent that the thrust bearing and the rotor side surface are not burned. More specifically (when A2 = A1 to A1> A2).
[0017]
In order to reduce the load P during operation, the inner area of the seal lip portion is set to A1> A2 in the present invention . The load on the bearing increases. Therefore, the value of (A2-A1) is such that the thrust bearing and the rotor side face are not burned out due to the relationship with the seal chamber pressure during stoppage.
[0018]
Here, the seal oil supply passage is supplied to the seal through a cooler from the oil separator, enters the rotor tooth groove through the oil passage from the seal chamber, and returns to the oil separator together with the discharge gas. Here, during operation, the pressure is reduced by the throttle immediately before sealing and the piping resistance between the oil separator and the seal oil supply port, and the seal chamber pressure is the pressure obtained by adding the flow resistance (seal chamber to rotor tooth groove) to the rotor tooth groove pressure. Become. Generally, this value is about suction pressure + 1 to 5 kgf / cm ² . The fluctuation of about 1 to 5 kgf / cm ² varies depending on the suction pressure and the amount of oil supplied. On the other hand, when the motor stops, the rotation of the rotor stops, so that the oil in the seal chamber is not discharged from the rotor, and the oil flow stops. For this reason, the piping resistance between the oil separator and the seal chamber is eliminated, and the pressure of the oil separator is properly applied to the seal chamber. Since the pressure of the oil separator after the stop is equalized, the pressure is reduced. However, the discharge pressure is still substantially maintained immediately before the rotation of the rotor is stopped, and the seal chamber pressure at the stop is substantially the discharge pressure.
[0019]
For this reason, in the equations (1) and (2) of the load P applied to the thrust bearing, the seal chamber pressure P1 at the time of stop is substantially the discharge pressure, and becomes the suction pressure + 1 to 5 kgf / cm ² during operation. Therefore, when it is desired to greatly aim to reduce the load P applied to the thrust bearing during operation, it is difficult to increase the value of (A2-A1) because of the relationship between the seal chamber pressure P1 during operation and stop. In this case, the value of (A1-A2) is set to such an extent that there is no problem at the time of stoppage (when A2 = A1 to A1> A2), and the seal chamber pressure P1 during operation is set by the throttle valve shown in FIG. It is the invention of claim 4 that enlarges and reduces the load P applied to the thrust bearing.
[0020]
From the following description, the object of the present invention can be achieved smoothly.
[0021]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, the present invention will be described in detail with reference to the embodiments shown in the drawings. However, the dimensions, materials, shapes, relative positions, and the like described in this embodiment are merely illustrative examples and not intended to limit the scope of the present invention unless otherwise specified.
1 and 2 are longitudinal sectional views of a male rotor portion of a screw compressor according to an embodiment of the present invention.
[0022]
1 and 2, 1 is a rotor casing, 5 is a male rotor housed in the rotor casing 1, 2 is a discharge side bearing housing, and 3 is a suction side bearing housing, which is located outside the rotor casing 1. Yes. Further, the discharge side seal housing 12A and the suction side seal housing 12B are connected to the outside of the bearing housings 2 and 3, respectively.
The discharge-side rotor shaft 5a and the suction-side rotor shaft 5b integrated with the male rotor 5 are rotatably supported by the bearing housings 2 and 3 assembled on both sides of the rotor casing 1 via radial bearings 7 and 8, respectively. Further, the discharge side rotor shaft 5a receives a thrust load via a thrust bearing 9 such as a pair of angular ball bearings. The discharge-side rotor shaft 5a is configured as a drive shaft to which an electric motor (not shown) is connected and the male rotor 5 is rotationally driven.
[0023]
Seal rotating members 10 and 11 of mechanical seals are fixed by shafts on the outer sides of the thrust bearings 9 and the radial bearings 7 and 8 of the rotor shafts 5a and 5b by a method not shown in the figure, while they are accommodated. Seal fixing members 10a and 11a are fitted on both end face sides of the seal rotation members 10 and 11 on the inner peripheral side of the discharge side seal housing 12A and the suction side seal housing 12B.
The lips of the seal fixing members 10a and 11a are in contact with the side surfaces of the seal rotating members 10 and 11, so that the contact surface pressure with the side surfaces of the seal rotating members 10 and 11 is appropriate via an elastic member (not shown). It is pressed and seals at the contact surface. The shaft end of the suction side rotor shaft 5b is open to the atmosphere.
Of course, the end of the shaft 5b may be covered with a cover provided with a small hole communicating with the outside air in order to prevent intrusion of dust or the like, but it is important that the shaft 5b is open to the atmosphere.
[0024]
Throttle valves 15a and 16a are provided in the oil supply passages 15 and 16 of the discharge side seal housing 12A and the suction side seal housing 12B facing the outer circumferences of the seal rotating members 10 and 11 of the mechanical seal, respectively, and the throttle valves 15a and 16a are provided via the throttle valves 15a and 16a. The lubricating oil is supplied from the oil separator 21 to the seal chambers 13 and 14 while lubricating the sliding portion of the mechanical seal, and the oil passages 19 'and 20' from the oil outlet passages 19 and 20 and the passage 25 in the casing 1 where they join. Is fed as rotor lubricating oil to the rotor chamber in which the rotor 5 is housed. The throttle valves 15a and 16a function suitably for adjusting the flow rate of the lubricating oil and controlling the hydraulic pressure immediately after the operation is stopped, and are not provided in the embodiment shown in FIG. The function of the present invention can be achieved without providing 16a. In addition, the oil which lubricated the bearing and the seal is discharged together with the gas fluid through an oil passage (not shown).
In FIG. 2, a throttle valve 26 is provided in an oil passage 25 ′ where the oil passages 19 ′ and 20 ′ join, and oil is supplied to the passage 25 in the casing 1 through the throttle valve 26. .
[0025]
FIG. 3 is a piping system diagram related to the screw compressor of the present invention. In FIG. 3, when supercritical operation is performed using CO ₂ as a refrigerant, ^the suction gas having a suction pressure of 20 to 50 kg / cm ² passes through the pipeline 5 . Te sucked from the suction port 17 to the screw compressor, ejection discharge pressure enters the oil separator 21 via line 36 from 100 kg / cm ² is compressed around the gas fluid (gas) discharge port 18, mixed The lubricating oil is separated and sent to the pipe line 27. Lubricating oil 24 separated from the discharge gas by the oil separator 21 is accumulated at the bottom of the oil separator 21, passes through a conduit 28, and is connected to the discharge side seal oil supply port 15 and the suction side seal oil supply port 16 to the seal chamber 13, 14. The number of oil supply ports to the seal chamber may be one, and both the seal chambers may communicate with each other by providing an oil passage in the housing and the casing. 22 and 23 are check valves on the discharge side and the suction side, respectively. The lubricating oil 24 is cooled through an oil cooler (not shown) and supplied to the screw compressor.
[0026]
Next, the operation will be described. During the operation of the screw compressor, a thrust force substantially equal to the discharge side pressure and the suction side pressure is applied to the rotor 5 in the suction side direction. On the other hand, the inside of the oil separator 21 is at a discharge pressure, and the lubricating oil fed from here to the seal chamber oil supply ports 15 and 16 enters the discharge side seal chamber 13 and the suction side seal chamber 14 from which oil is supplied. It spouts into a rotor chamber through the paths 19 and 20. The pressure in the seal chambers 13 and 14 is the pressure obtained by adding the flow resistance (seal chamber to rotor tooth groove) to the rotor tooth groove pressure. Generally, this value is about suction pressure + 1 to 5 kgf / cm ² .
[0027]
The thrust force P applied to the thrust bearing is calculated from P = (A2−A1) from the lip inner diameter areas A1 and A2 of the seal fixing members 10a and 11a, the discharge side seal chamber pressure P1, the suction side seal pressure P2, and the gas load F.・ P1 + F.
Here, when the load direction is + in the suction side direction, F is always +. If A2 = A1, the load applied to the thrust bearing at the time of stop is P = 0, and only F acts on the thrust bearing during operation. Here, in order to reduce the load P applied to the thrust bearing during operation, the inner diameter area of the seal lip is set to A1> A2. However, if A1> A2, the load applied to the thrust bearing at the time of stopping increases, so the set value of (A2-A1) must be determined to the extent that the thrust bearing and the rotor side face are not burned.
[0028]
In the case of a high discharge pressure, the pressure in the seal chambers 13 and 14 also increases, and the amount of lubricating oil ejected from the seal chambers 13 and 14 through the oil passages 19 and 20 may be excessive. In that case, the amount of lubricating oil can be adjusted by adjusting the opening of the throttle valve 26. That is, the seal chamber pressure P1 during operation is increased by the throttle valve 26 shown in FIG. 2 to reduce the load P applied to the thrust bearing.
FIG. 4 is a piping system diagram of a second embodiment relating to the screw compressor of the present invention. The same reference numerals are given to the same components as those in FIG. The difference from FIG. 2 is that throttle valves 15 a and 16 a whose opening degree can be adjusted are arranged in an oil supply line 28 from the oil separator 21 to the oil supply ports 15 and 16. The narrowed Riben 15a during operation, 16a is pressure of the sealing chamber 13, 14 in the fully open or normal opening is approximately the suction pressure + 1~5kgt / cm ^2. Immediately before the operation is stopped, the throttle valves 15a and 16a and all the lines for supplying oil to the compressor are closed or greatly reduced to reduce the pressure in the seal chambers 13 and 14, and then the operation is stopped. Therefore, since the pressure in the seal chambers 13 and 14 immediately after the operation is stopped is lower than the discharge pressure during operation, the discharge side thrust force due to the pressure in the seal chambers 13 and 14 is smaller than that during operation, and the operation is stopped. It is possible to prevent thrust bearing damage and rotor end face burnout due to excessive discharge side direction thrust force. Note that the time from the throttle valve operation to the operation stop operation is extremely short, and the throttle valve operation and the operation stop operation are automatically performed at an appropriate time interval by the stop operation.
[0030]
【The invention's effect】
As described above, according to the present invention, in the screw compressor used for high pressure compression, it is possible to suppress the occurrence of excessive reverse thrust load of the thrust bearing that is suddenly reversed in the reverse direction when the operation is stopped, and when the operation is stopped. It is possible to eliminate the generated thrust bearing damage and the burning of the rotor end face.
[Brief description of the drawings]
FIG. 1 is a longitudinal sectional view of a male rotor portion of a screw compressor according to a first embodiment of the present invention.
FIG. 2 is a longitudinal sectional view of a male rotor portion of a screw compressor according to another embodiment of the present invention.
FIG. 3 is a piping system diagram according to a first embodiment of the screw compressor of FIGS. 1 and 2;
4 is a piping system diagram according to a second embodiment of the screw compressor of FIGS. 1 and 2. FIG.
FIG. 5 is a longitudinal sectional view of a conventional screw compressor.
[Explanation of sign]
DESCRIPTION OF SYMBOLS 1 Rotor casing 2 Bearing housing 3 Bearing housing 4 Seal housing 5 Male rotor 6 Female rotor 7 Radial bearing 9 Thrust bearing 10 Seal rotation member 10a Seal fixing member 11 Seal rotation member 12 Lock nut 13 Discharge side seal chamber 14 Suction side seal chamber 21 Oil separators 15a, 16a, 26 Throttle valve 28 Oil supply line

Claims (5)

A rotor casing that houses a rotor chamber in which a male rotor and a female rotor that have shaft portions on both sides and are engaged with each other are rotatably supported in parallel, and screw compression that includes suction and discharge ports on both ends of the rotor of the casing In a screw compression apparatus comprising a machine main body and an oil separator that separates lubricating oil mixed in the compressed fluid discharged from the discharge port,
The male rotor serving as a drive shaft is provided with a mechanical seal portion including a seal rotating member fixed to the shaft and a seal fixing member fixed to the casing on the discharge side and the suction side of the male rotor shaft. While opening both ends of the shaft to the outside air,
Forming a suction-side seal chamber for applying a supply hydraulic pressure from the oil separator on the surface opposite to the outside air side of the suction-side seal rotating member;
A discharge-side seal that sets the discharge-side seal rotating member to have a larger diameter than the suction-side seal rotating member and applies the supply hydraulic pressure from the oil separator to the surface opposite to the outside air side of the discharge-side seal rotating member Forming a chamber,
Reducing the thrust force in the discharge side direction due to the pressure difference between the pressure in the rotor chamber and the atmospheric pressure generated when the operation is stopped with the outside of the seal rotating members of the suction side seal chamber and the discharge side seal chamber open to the atmosphere ,
A high-pressure screw compressor configured to relieve a thrust force in a suction side direction due to a differential pressure between a discharge pressure and a suction pressure generated during operation by the hydraulic pressure in both the seal chambers.   2. The high pressure according to claim 1, further comprising means for adjusting a hydraulic pressure to reduce the thrust force applied to the discharge side when the operation is stopped by restricting the supply of the hydraulic pressure to each seal chamber immediately before the operation is stopped. Screw compression device.   3. The high-pressure screw compressor according to claim 2, wherein the means for adjusting the hydraulic pressure is a throttle portion provided in the middle of a passage where the hydraulic pressure separated by the oil separator is applied to each seal chamber. apparatus.   A throttle valve is provided in a supply path for supplying lubricating oil from the discharge-side seal chamber to the rotor chamber between the casing and the rotor, and the pressure in the discharge-side seal chamber is adjusted by the throttle valve during operation of the screw compressor. 2. The high-pressure screw compressor according to claim 1, wherein a thrust force applied in the suction side direction can be reduced.   2. The high-pressure screw compressor according to claim 1, wherein the mechanical seal portion is provided on the outer side of a bearing portion that supports shaft portions on the discharge side and the suction side of the male rotor shaft.

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