CN111295518A

CN111295518A - Liquid-cooled screw compressor

Info

Publication number: CN111295518A
Application number: CN201880072460.1A
Authority: CN
Inventors: 藤泽亮; 田中孝二; 野口透; 坂口广宣
Original assignee: Kobe Steel Ltd
Current assignee: Shengang Compressor Co ltd
Priority date: 2017-11-09
Filing date: 2018-10-22
Publication date: 2020-06-16
Anticipated expiration: 2038-10-22
Also published as: CN111295518B; JP2019085969A; WO2019093109A1; JP6836492B2

Abstract

The invention relates to an oil-cooled screw compressor (1), comprising: a screw rotor (40) including a male rotor (50) and a female rotor (60) meshing with the male rotor (50); a rotor case (10) that houses the screw rotor (40); and a plurality of oil supply ports (13) which are provided in the rotor case (10) and are configured to supply oil to at least two tooth space spaces of the screw rotor (40).

Description

Liquid-cooled screw compressor

Technical Field

The present invention relates to a liquid-cooled screw compressor.

Background

In an oil-cooled screw compressor, which is one type of liquid-cooled screw compressors, for example, patent document 1 discloses a technique for promoting heat exchange between gas and oil during compression by studying the arrangement of an injection nozzle (oil supply port). In the oil-cooled screw compressor of patent document 1, the injection direction from the injection nozzle into the compression chamber (rotor chamber) is directed in the direction opposite to the rotation direction of the screw rotor. This ensures a long time for the oil to fly in the gas in the compression chamber, and promotes heat exchange between the oil and the gas.

Documents of the prior art

Patent document

Patent document 1: japanese patent laid-open No. 9-151870.

Disclosure of Invention

Problems to be solved by the invention

The oil-cooled screw compressor compresses gas by rotating the screw rotors at high speed, and therefore it is considered that the time for which the oil is in contact with the air before the oil is in contact with the rotor casing and the screw rotors increases only slightly even if the injection direction is changed as described above. Therefore, the heat exchange performance is expected to be slightly improved, and there is room for improvement in order to further improve the heat exchange performance. Further, no special investigation has been made for the improvement of the sealing performance by oil between the male rotor, the female rotor, and the rotor case. In particular, in the oil-cooled screw compressor of patent document 1, oil is supplied to the low-pressure portion, but there is a possibility that the oil is insufficient in the high-pressure portion. Further, if the amount of oil delivered to the high-pressure portion is increased by increasing the amount of oil supplied to the low-pressure portion, the oil is excessively supplied to the low-pressure portion, and power loss occurs due to the excessive oil agitation.

The invention aims to improve heat exchange performance and sealing performance by preventing liquid (oil) offset in a liquid-cooled screw compressor, particularly an oil-cooled screw compressor.

Means for solving the problems

The invention provides an oil-cooled screw compressor, comprising: a screw rotor including a male rotor and a female rotor meshing with the male rotor; a rotor case that houses the screw rotor; and a plurality of liquid supply ports arranged in the rotor case so as to supply liquid into at least two tooth space spaces of the screw rotor.

According to this configuration, the liquid supply port is disposed in the rotor case so as to supply liquid into at least two tooth space spaces of the screw rotor, and therefore, the offset of the liquid can be prevented. Here, the cogging space means a communication space defined by the rotor case, the teeth of the male rotor, and the teeth of the female rotor. The liquid can be prevented from being biased, so that the heat exchange between the liquid and the compressed gas can be promoted and the heat exchange performance can be improved, and the liquid-based sealing performance between the male rotor and the female rotor and the rotor housing can be improved. Therefore, the compression efficiency is improved, so that the energy saving performance can be improved.

The liquid supply amount adjusting mechanism may be further provided to decrease the amount of liquid supplied from the liquid supply port as the liquid flows from the low pressure side to the high pressure side.

According to this structure, the offset of the liquid can be further prevented. Since the volume of the gas in the spline space decreases as it is compressed, the amount of liquid supply required to travel from the low-pressure portion to the high-pressure portion decreases. Therefore, the amount of liquid supplied can be reduced by the liquid supply amount adjustment mechanism as the liquid moves from the low pressure portion to the high pressure portion, and a desired amount of liquid can be supplied to an appropriate portion. This prevents excessive liquid supply to the low-pressure portion, and thus reduces power loss due to excessive liquid agitation, thereby improving energy saving performance.

The plurality of liquid supply ports may be arranged on a straight line, and a straight liquid supply pipe may be provided to connect the liquid supply ports arranged on the straight line.

According to this configuration, since the liquid supply pipe is linear, the liquid supply pipe can be prevented from being complicated in shape and the amount of work for machining the liquid supply pipe can be reduced. The liquid supply ports need not be all provided on a straight line, and there may be a plurality of liquid supply ports arranged on a straight line, or liquid supply ports arranged outside the straight line.

At least two of the plurality of liquid supply ports may be provided on one side of the male rotor or the female rotor, and a distance between farthest points of the adjacent liquid supply ports provided on the one side may be smaller than a width of the tooth groove on the one side of the male rotor or the female rotor in a direction in which the screw rotor extends.

According to this structure, at least two liquid supply ports can be arranged in one tooth space. Therefore, the shortage of liquid in the tooth grooves can be suppressed. Therefore, the cooling performance and the sealing performance by a sufficient amount of liquid are improved, and the compression efficiency can be improved.

At least two of the plurality of liquid supply ports may be provided on one side of the male rotor or the female rotor, and a distance between nearest points of adjacent liquid supply ports provided on the one side may be larger than a width of the tooth space on the one side or a distance between farthest points of adjacent liquid supply ports provided on the one side may be smaller than a width of the tooth space on the one side in a direction in which the screw rotor extends.

According to this structure, noise can be reduced. If the distance between the closest points of the adjacent liquid supply ports is smaller than the width of the tooth grooves on the side where the liquid supply ports are provided in the male rotor or the female rotor, and the distance between the farthest points of the adjacent liquid supply ports is larger than the width of the tooth grooves on the side where the liquid supply ports are provided in the male rotor or the female rotor, the tooth tips become located near the liquid supply ports at the same time. Since the pressure in the liquid supply ports increases sharply due to centrifugal force when the tooth tip passes through, if a sharp pressure increase is caused simultaneously in two adjacent liquid supply ports, the pressure in the liquid supply pipe rises sharply. A rapid rise in the pressure in the liquid supply pipe causes pulsation, which causes noise. Therefore, by providing the liquid supply port while avoiding the above-described situation, noise can be reduced.

The rotor housing may have a suction port at a position corresponding to an end of the screw rotor in a direction in which the screw rotor extends, and a portion of the plurality of liquid supply ports closest to the liquid supply port may be separated from the suction port by a tooth space width of the male rotor or the female rotor on a side of the liquid supply port closest to the suction port.

According to this configuration, even if the screw rotor rotates, the liquid supply port (the lowest pressure side) closest to the suction port and the suction port are not fluidly connected. Therefore, the liquid can be prevented from leaking to the suction port, the suction heating can be prevented, and the volume efficiency can be prevented from being lowered.

The rotor housing may have a discharge port at a position corresponding to an end of the screw rotor in a direction in which the screw rotor extends, and a position of the liquid supply port closest to the discharge port among the plurality of liquid supply ports may be separated from the discharge port by a width of a tooth space on the side of the liquid supply port closest to the discharge port in the male rotor or the female rotor.

According to this configuration, even if the screw rotor rotates, the liquid supply port (on the highest pressure side) closest to the discharge port and the discharge port are not fluidly connected. Therefore, the liquid can be prevented from flowing backward from the discharge port into the liquid supply port, the volumetric efficiency can be prevented from being lowered, and the power loss due to recompression can be prevented.

Effects of the invention

According to the present invention, in the liquid-cooled screw compressor, since liquid is supplied into at least two tooth space spaces of the screw rotor, the liquid can be prevented from being biased, and the heat exchange performance and the sealing performance can be improved.

Drawings

FIG. 1 is a partial schematic configuration diagram of an oil-cooled screw compressor according to a first embodiment of the present invention,

figure 2 is a schematic cross-sectional view of the rotor housing along line II-II of figure 1,

figure 3 is a schematic cross-sectional view showing the position of an oil supply port in the rotor case,

figure 4 is a schematic cross-sectional view showing the position of an oil supply port in the rotor case,

fig. 5 is a sectional view showing the arrangement of a rotor casing of the oil-cooled screw compressor of the first modification,

fig. 6 is a sectional view showing the arrangement of a rotor casing of an oil-cooled screw compressor of a second modification,

FIG. 7 is a schematic sectional view showing the position of an oil supply port in a rotor case of an oil-cooled screw compressor according to a second embodiment,

fig. 8 is a schematic cross-sectional view showing the position of an oil supply port in a rotor case of an oil-cooled screw compressor according to a third embodiment.

Description of the reference numerals

1 compressor (oil cooling type screw compressor)

10 rotor case

10a suction inlet

10b discharge port

11. 12 partition wall

13 oil supply port

13a to 13c oil supply port (oil supply amount adjusting mechanism)

13A to 13C flow control valve (fuel supply amount control mechanism)

14 point of point

14a suction side tip point

14b discharge side tip point

15 oil supply pipe

15a, 15b, 15c (oil supply amount adjusting mechanism)

20. 21 bearing shell

30 rotor chamber

31 male rotor chamber

32 female rotor chamber

33. 34 bearing chamber

40 screw rotor

50 male rotor

51. 52-shaft member

53. 54 bearing.

60 female rotor

61. 62 shaft member

63. 64 bearings.

Detailed Description

Embodiments of the present invention will be described below with reference to the drawings. As a device according to an embodiment of the present invention, an oil-cooled screw compressor using oil as a liquid to be supplied into a rotor case is shown. Therefore, the "oil" and "liquid" may be interchanged below.

(first embodiment)

Fig. 1 is a partial schematic configuration diagram of an oil-cooled screw compressor 1 according to a first embodiment of the present invention. Hereinafter, the oil-cooled screw compressor 1 is also simply referred to as the compressor 1. Fig. 1 shows a portion of a compressor 1 particularly relating to a compression mechanism. The compressor 1 sucks air from the outside, compresses the air in the inside, and discharges the air. The air discharged from the compressor 1 is supplied to a supply destination through a pipe not shown.

The compressor 1 includes a rotor housing 10 and

bearing housings

20 and 21. In the present embodiment, the rotor housing 10 and the bearing

housings

20 and 21 are integrated. The rotor housing 10 is arranged between the two bearing

housings

20, 21. The rotor housing 10 defines a rotor chamber 30 inside, and the two bearing

housings

20, 21 define bearing

chambers

33, 34 inside, respectively. Rotor chamber 30 and bearing chamber 33 are partitioned by partition wall 11, and rotor chamber 30 and bearing chamber 34 are partitioned by partition wall 12. Both

partition walls

11, 12 are part of the rotor housing 10.

In the rotor housing 10, a male rotor 50 and a female rotor 60 that meshes with the male rotor 50 and has more teeth than the male rotor 50 are disposed. That is, the male rotor 50 and the female rotor 60 constitute the screw rotor 40. Although not shown in detail, in the present embodiment, the male rotor 50 has 4 teeth and the female rotor 60 has 6 teeth, for example.

Fig. 2 is a schematic sectional view of the rotor case 10 taken along line II-II of fig. 1. The rotor housing 10 defines a male rotor chamber 31 that houses the male rotor 50 and a female rotor chamber 32 that houses the female rotor 60. The rotor chamber 30 is a space formed by joining the male rotor chamber 31 and the female rotor chamber 32. The rotor case 10 has a shape in which two cylinders are connected at the side, in other words, the male rotor chamber 31 and the female rotor chamber 32 are both cylindrical spaces, communicating with each other.

Fig. 2 is also a cross-sectional view viewed from the direction of the rotation axis of the female rotor 60 (see fig. 1). In the present embodiment, the rotation axis of the female rotor 60 and the rotation axis of the male rotor 50 extend horizontally in parallel with each other, and the male rotor chamber 31 and the female rotor chamber 32 extend in the same direction. In the sectional view of fig. 2, the male rotor chamber 31 and the female rotor chamber 32 are connected by the suction-side tip point 14a and the discharge-side tip point 14 b. The lowermost point P3 of the male rotor chamber 31 is located below the discharge-side tip point 14b connecting the male rotor chamber 31 and the female rotor chamber 32.

As shown in fig. 1, a shaft member 51 forming a rotation shaft of the male rotor 50 extends from one end of the male rotor 50. The shaft member 51 extends from the male rotor chamber 31 to the bearing chamber 33 through the partition wall 11, and is rotatably supported by the bearing 54 in the bearing chamber 33. Further, a shaft member 61 forming a rotation shaft of the female rotor 60 extends from one end of the female rotor 60. The shaft member 61 extends from the female rotor chamber 32 to the bearing chamber 33 through the partition wall 11, and is rotatably supported by a bearing 63 in the bearing chamber 33.

A shaft member 52 forming a rotation shaft of the male rotor 50 extends from the other end of the male rotor 50. The shaft member 52 extends from the male rotor chamber 31 to the bearing chamber 34 through the partition wall 12, and is rotatably supported by a bearing 54 in the bearing chamber 34. Further, a shaft member 62 forming a rotation shaft of the female rotor 60 extends from the other end of the female rotor 60. The shaft member 62 extends from the female rotor chamber 32 to the bearing chamber 34 through the partition wall 12, and is rotatably supported by a bearing 64 in the bearing chamber 34. In particular, the shaft member 52 of the male rotor 50 extends to and is mechanically connected to a motor, not shown. Therefore, the male rotor 50 is rotationally driven by the motor, rotational power is transmitted from the male rotor 50 to the female rotor 60, and the male rotor 50 and the female rotor 60 are engaged with each other to rotate, thereby compressing air. In fig. 1, the right side is a suction side, and the left side is a discharge side. Therefore, if the male rotor 50 and the female rotor 60 rotate, air is sucked from the bearing chamber 34 side and discharged to the bearing chamber 33 side in the rotor chamber 30.

As shown in fig. 2, the rotor case 10 is provided with the oil supply port 13 on the female rotor chamber 32 side. Three virtual line segments S1-S3 are defined to define the detailed position of the fuel fill inlet 13. A first virtual line segment S1 is defined that connects the rotation center point P1 of the female rotor 60 and the discharge-side tip point 14 b. A second virtual line segment S2 is defined in which the first virtual line segment S1 is rotated by a first predetermined angle θ 1 in a direction away from the male rotor 50 about the rotational center point P1 of the female rotor 60. Here, the first predetermined angle θ 1 is an angle to the extent that concentration of oil to the meshing position (approximately aligned with the discharge-side tip point 14 b) can be prevented, and is determined according to the shape of the male rotor 50 and the female rotor 60, the shape of the rotor case 10, the type of oil, and the like. A first center line L1 including a rotation center point P2 of the male rotor 50 and a rotation center point P2 of the female rotor 60 is defined. The first center line L1 is a horizontal line. Further, a second center line L2 perpendicular to the first center line L1 and passing through the rotational center point P1 of the female rotor 60 is defined. A third virtual line segment S3 is defined that is rotated by a second predetermined angle θ 2 from the second center line L2 in a direction away from the discharge-side tip point 14b about the rotational center point P1 of the female rotor 60. Here, the second predetermined angle θ 2 is an angle to prevent the shortage of oil at the meshing position, and is determined according to the shape of the male rotor 50 and the female rotor 60, the shape of the rotor case 10, the type of oil, and the like.

The fuel fill inlet 13 is provided in a range excluding the range from the first virtual line segment S1 to the second virtual line segment S2, and more specifically, preferably in a range from the second virtual line segment S2 to the third virtual line segment S3. More specifically, a part of the oil supply port 13 is preferably provided in a range from the second virtual line segment S2 to the third virtual line segment S3. Here, the first predetermined angle θ 1 is preferably, for example, about 30 degrees or more. The second predetermined angle θ 2 is, for example, about 1/4 or less of the angle corresponding to one tooth of the female rotor 60, and is preferably about 15 degrees or less since the female rotor 60 having 6 teeth is used in the present embodiment. In the present embodiment, the fuel fill inlet 13 is provided in the range from the second virtual line segment S2 to the third virtual line segment S3, specifically, on the second center line L2.

By providing the oil supply port 13 in the above range, oil can be prevented from being concentrated at the meshing position of the male rotor 50 and the female rotor 60. In general, in the oil-cooled screw compressor 1, oil tends to concentrate at the meshing position of the male rotor 50 and the female rotor 60 due to the rotation of the male rotor 50 and the female rotor 60. If the oil supply port 13 is provided in the range from the first virtual line segment S1 to the second virtual line segment S2, oil is supplied to the vicinity of the meshing position between the male rotor 50 and the female rotor 60, and therefore, oil is concentrated at the meshing position, which may cause an excessive stirring loss of oil and deterioration of compression efficiency. However, in the configuration of the present embodiment, since the oil fill port 13 is provided at a position (a range excluding the range from the first virtual line segment S1 to the second virtual line segment S2) that is apart from the meshing position to some extent, concentration of oil at the meshing position can be prevented, and deterioration of compression efficiency can be prevented.

Further, by providing the oil supply port 13 in the range from the second virtual line segment S2 to the third virtual line segment S3, the sealability at the meshing position of the male rotor 50 and the female rotor 60 can be ensured. If the oil supply port 13 is disposed at a position greatly distant from the meshing position, there is a possibility that oil cannot be sufficiently supplied to the meshing position of the male rotor 50 and the female rotor 60. In this case, air may leak out at the meshing position of the male rotor 50 and the female rotor 60, and the compression efficiency may deteriorate. However, in the configuration of the present embodiment, since the oil fill port 13 is provided at a position (in the range from the second virtual line segment S2 to the third virtual line segment S3) close to the meshing position to some extent, the oil can be sufficiently supplied to the meshing position. Therefore, the sealability at the meshing position can be ensured, and the deterioration of the compression efficiency can be prevented.

Fig. 3 and 4 are schematic cross-sectional views showing the position and size of the oil supply port 13 in the rotor case 10. In the present embodiment, a plurality of (three in fig. 3 and 4)

fuel fill ports

13a, 13b, and 13c are arranged on a straight line so as to be connected by a straight line fuel supply pipe 15. The oil supply pipe 15 extends parallel to the rotation axis of the female rotor 60.

The oil supply ports 13a to 13c are preferably arranged to supply oil into at least two tooth space spaces of the screw rotor 40. The cogging space refers to a communication space defined by the rotor housing 10, the teeth of the male rotor 50, and the teeth of the female rotor 60. For example, one gullet space is represented by a hatched portion in FIG. 3. In the present embodiment, the distance D1 between the farthest points of the fuel fill port 13a and the fuel fill port 13c (the distance D1 between the fuel fill port 13a and the farthest portion of the fuel fill port 13 c) is greater than the tooth space width D1 of the female rotor 60 (D1 > D1). Thus, the oil fill port 13a and the oil fill port 13c are located in different slot spaces, and therefore oil can be supplied to at least two slot spaces.

The distance D2 between the farthest points of the oil fill port 13a and the oil fill port 13b adjacent to each other (the distance D2 between the oil fill port 13a and the farthest portion of the oil fill port 13 b) is smaller than the tooth space width D1 of the female rotor 60 (D2 < D1). Similarly, the distance D3 between the farthest points of the oil supply port 13b and the oil supply port 13c (the distance D3 between the farthest points of the oil supply port 13b and the oil supply port 13 c) adjacent to each other is smaller than the tooth space width D1 of the female rotor 60 (D3 < D1). According to these dimensional relationships, when the female rotor 60 rotates and the tooth tips thereof move, at least two

fuel supply ports

13a, 13b or

fuel supply ports

13b, 13c are arranged in one tooth groove. Further, with this arrangement, as will be described later, the tooth tips of the female rotor 60 are not simultaneously positioned at the two adjacent fuel fill

ports

13a, 13b or the two adjacent fuel fill

ports

13b, 13c, and therefore, as will be described later, a rapid increase in the pressure in the fuel supply pipe 15 can be prevented.

The radii r 1-r 3 of the three oil supply ports 13 a-13 c are set so as to reduce the amount of oil supply as the oil supply port moves from the low pressure side to the high pressure side. Specifically, radius r1 is the largest, radius r2 is the second largest, and radius r3 is the smallest. In the present embodiment, for example, the radius r1 is about 1.5 times the radius r2, and the radius r2 is about 1.5 times the radius r 3. In the present embodiment, the three fuel supply ports 13a to 13c constitute a fuel supply amount adjusting mechanism.

As shown in fig. 3 and 4, the rotor case 10 has a suction port 10a at one end side (upper side in fig. 3, right side in fig. 4) of the screw rotor 40 in the rotation axis direction. The suction port 10a is provided in the upper portion of the rotor case 10 and extends around the male rotor 50 and the female rotor 60. Further, the position of the oil supply port 13a closest to the suction port 10a among the plurality of oil supply ports 13a to 13c in the rotation axis direction is separated from the suction port 10a by a tooth groove width D1 of the female rotor 60 or more. That is, the distance Δ 1 from the oil supply port 13a to the suction port 10a is equal to or greater than the tooth groove width D1 of the female rotor 60(Δ 1 ≧ D1).

Further, the rotor case 10 has a discharge port 10b on the other end side (lower side in fig. 3, left side in fig. 4) in the rotation axis direction. The discharge port 10b is provided at a lower portion of the rotor case 10. The position of the oil fill port 13c closest to the discharge port 10b among the plurality of oil fill ports 13a to 13c in the rotation axis direction is separated from the discharge port 10b by a tooth groove width D1 of the female rotor 60 or more. That is, the distance Δ 2 from the oil supply port 13c to the discharge port 10b is equal to or greater than the tooth groove width D1 of the female rotor 60(Δ 2 ≧ D1).

The operational effects of the compressor 1 of the present embodiment will be described below.

According to the present embodiment, the oil supply port 13 is disposed in the rotor case 10 so as to supply oil into at least two slot spaces of the screw rotor 40, and therefore, the offset of the oil can be prevented. By being able to prevent the offset of oil, the heat exchange of oil with compressed gas can be promoted and the heat exchange performance can be improved, and also the oil-based sealing performance between the male and

female rotors

50 and 60 and the rotor case 10 can be improved. Therefore, the compression efficiency is improved, so that the energy saving performance can be improved.

Further, according to the present embodiment, the offset of the oil in the rotor case 10 can be further prevented. Since the volume of the gas in the spline space decreases as it is compressed, the amount of oil supply required to travel from the low-pressure portion to the high-pressure portion decreases. Therefore, by gradually reducing the size of the oil supply ports 13a to 13c and reducing the amount of oil supply as the pressure goes from the low pressure portion to the high pressure portion, a required amount of oil can be supplied to an appropriate portion. Accordingly, excessive oil supply to the low-pressure portion can be prevented, and thus the power loss due to excessive oil agitation can be reduced, thereby improving the energy saving performance.

Further, according to the present embodiment, since the supply pipe 15 is linear, the shape of the supply pipe 15 can be prevented from being complicated, and the amount of work for machining the supply pipe 15 can be reduced. It is not necessary that all of the fuel fill ports 13 be provided on a straight line, and there may be a plurality of fuel fill ports 13 arranged on a straight line, or fuel fill ports 13 arranged outside the straight line.

Further, according to the present embodiment, since the farthest point distances d2 and d3 of the adjacent fuel fill ports 13 are defined as described above, at least two fuel fill ports 13 can be arranged in one tooth groove. Therefore, the oil shortage in the tooth grooves can be suppressed. Therefore, the cooling performance and the sealing performance by a sufficient amount of oil are improved, and the compression efficiency can be improved.

Further, according to the present embodiment, noise can be reduced. This is illustrated with reference to fig. 3. If the nearest point distances D4, D5 (the distance between the nearest portions of adjacent oil supply ports) of the adjacent oil supply ports 13 are smaller than the tooth space width D1 of the female rotor 60 and the farthest point distances D2, D3 are larger than the tooth space width D1(D4, D5 < D1 < D2, D3), the tooth tips of the female rotor 60 are located at the adjacent

oil supply ports

13a, 13b or 13b, 13c at the same time. Since the pressure in the fuel fill port 13 increases rapidly due to centrifugal force when the tooth tip passes through, if rapid pressure increases occur simultaneously in two adjacent fuel fill ports, the pressure in the fuel supply pipe 15 increases rapidly. The rapid rise in the pressure in the supply pipe 15 causes pulsation, which causes noise. Therefore, by providing the fuel fill port 13 while avoiding the above-described situation, noise can be reduced.

In the present embodiment, even when the screw rotor 40 rotates, the slot space where the oil supply port 13a (on the lowest pressure side) closest to the suction port 10a opens and the suction port 10a are defined with the tip of the screw rotor 40 interposed therebetween. Therefore, the oil fill port 13a and the suction port 10a do not directly communicate with each other. That is, according to the present embodiment, even if the screw rotor 40 rotates, the oil supply port 13a (on the lowest pressure side) closest to the suction port 10a is not fluidly connected to the suction port 10 a. Therefore, oil leakage to the suction port, suction heating, and a reduction in volumetric efficiency can be prevented.

Further, according to the present embodiment, even when the screw rotor 40 rotates, the tooth space in which the oil supply port 13c (on the highest pressure side) closest to the discharge port 10b opens and the discharge port 10b are defined with the tooth tips of the screw rotor 40 interposed therebetween. Therefore, the filler inlet 13c and the discharge port 10b do not directly communicate with each other. That is, according to the present embodiment, even if the screw rotor 40 rotates, the oil supply port 13c (on the highest pressure side) closest to the discharge port 10b is not fluidly connected to the discharge port 10 b. Therefore, the oil can be prevented from flowing backward from the discharge port into the oil supply port, the volumetric efficiency can be prevented from being lowered, and the power loss due to recompression can be prevented.

(first modification)

Fig. 5 is a sectional view of the rotor housing 10 of the compressor 1 according to the first modification, and corresponds to fig. 2. In the compressor 1 of the present modification, the rotor case 10 is disposed with the first center line L1 inclined from the horizontal line HL.

In the present modification, the rotation shaft CL1 of the female rotor 60 and the rotation shaft CL2 of the male rotor 50 are not arranged in the horizontal plane, and the rotation shaft CL1 of the female rotor 60 is arranged below the rotation shaft CL2 of the male rotor 50. Specifically, the first center line L1 forms about 30 degrees with the horizontal line HL, for example. Therefore, the female rotor chamber 32 is disposed below the male rotor chamber 31.

In the present embodiment, as described above, the lowermost point P3 of the male rotor chamber 31 is located below the discharge-side tip point 14b connecting the male rotor chamber 31 and the female rotor chamber 32.

(second modification)

Fig. 6 is a sectional view of the rotor housing 10 of the compressor 1 according to the second modification, and corresponds to fig. 2. In the compressor 1 of the present modification, the rotor case 10 is disposed in a state where the first center line L1 is vertical.

In the present modification, the rotation shaft CL1 of the female rotor 60 and the rotation shaft CL2 of the male rotor 50 are not arranged in the horizontal plane, and the rotation shaft CL2 of the male rotor 50 is arranged slightly below the rotation shaft CL1 of the female rotor 60, so that the entire male rotor chamber 31 is arranged below the entire female rotor chamber 32.

In the present modification, as in the first modification, the lowermost point P3 of the male rotor chamber 31 is located below the discharge-side tip point 14b connecting the male rotor chamber 31 and the female rotor chamber 32.

(second embodiment)

Fig. 7 is a sectional view of the rotor housing 10 of the compressor 1 according to the second embodiment, and corresponds to fig. 4 of the first embodiment. The compressor 1 of the present embodiment is the same as the compressor 1 of the first embodiment except for the configuration of the oil supply port 13. Therefore, the same components as those in the first embodiment are denoted by the same reference numerals, and description thereof is omitted.

In the present embodiment, the diameters of the three fuel supply ports 13a to 13c are the same, unlike the first embodiment. Flow rate regulating valves 13A to 13C are mounted to the oil supply ports 13A to 13C. The flow rate control valves 13A to 13C are set so as to reduce the amount of fuel supplied from the fuel supply ports 13A to 13C as the fuel advances from the low pressure side (right side in FIG. 7) to the high pressure side (left side in FIG. 7). Specifically, the allowable flow rate of the flow rate adjustment valve 13A is maximized, the allowable flow rate of the flow rate adjustment valve 13B is maximized, and the allowable flow rate of the flow rate adjustment valve 13C is minimized. In the present embodiment, the flow rate adjustment valves 13A to 13C constitute an oil supply amount adjustment mechanism.

(third embodiment)

Fig. 8 is a sectional view of the rotor housing 10 of the compressor 1 according to the third embodiment, and corresponds to fig. 4 of the first embodiment. The compressor 1 of the present embodiment is the same as the compressor 1 of the first embodiment except for the configuration of the oil supply port 13. Therefore, the same components as those in the first embodiment are denoted by the same reference numerals, and description thereof is omitted.

In the present embodiment, the diameters of the three fuel supply ports 13a to 13c are the same, unlike the first embodiment. Three oil supply pipes 15a to 15c are connected to the oil supply ports 13a to 13c, respectively. The fuel supply amount from each of the fuel supply ports 13a to 13c is set to decrease as the vehicle travels from the low pressure side (left side in fig. 8) to the high pressure side (right side in fig. 8). Specifically, the thicknesses of the supply pipes 15a to 15c are different from each other, and the supply pipe 15a is the thickest, the supply pipe 15b is the second thickest, and the supply pipe 15c is the thinnest. In the present embodiment, the fuel supply pipes 15a to 15c constitute fuel supply amount adjusting means.

While the present invention has been described with reference to the specific embodiments, the present invention is not limited to the above embodiments, and various modifications can be made within the scope of the present invention.

In the above embodiment, all the fuel fill ports 13 are arranged on a straight line, but the arrangement of the fuel fill ports 13 is not limited to this. For example, the fuel fill inlet 13 may be partially arranged on a straight line and the fuel fill inlet 13 may be provided without being arranged on a straight line. The number and size of the fuel-filling ports 13 are not particularly limited.

In the above embodiments, the arrangement of the fuel supply port 13 provided on the female rotor 60 side can be similarly applied to the male rotor 50 side, and vice versa. That is, the arrangement structure of the fuel fill inlet 13 in each of the above embodiments may be adopted for either one or both of the male rotor 50 and the female rotor 60 without distinguishing the male rotor 50 from the female rotor 60.

As described above, the oil-cooled compressor using oil as the liquid supplied into the rotor case is shown as the device according to the embodiment of the present invention. However, the present invention can also be applied to a liquid-cooled compressor other than the oil-cooled compressor. For example, the present invention can also be applied to a water jet compressor using water as the liquid to be supplied into the rotor casing.

Claims

1. A liquid-cooled screw compressor comprising:

a screw rotor including a male rotor and a female rotor meshing with the male rotor;

a rotor case that houses the screw rotor; and

and a plurality of liquid supply ports arranged in the rotor case so as to supply liquid into at least two tooth space spaces of the screw rotor.

2. The liquid cooled screw compressor of claim 1,

and a liquid supply amount adjusting mechanism for decreasing the amount of liquid supplied from the liquid supply port as the liquid flows from the low pressure side to the high pressure side.

3. The liquid cooled screw compressor of claim 2,

the plurality of liquid supply ports are arranged on a straight line,

the liquid-cooled screw compressor further includes a linear liquid supply pipe connecting the liquid supply ports arranged on a straight line.

4. Liquid cooled screw compressor according to any of claims 1 to 3,

at least two of the plurality of liquid supply ports are provided on one side of the male rotor or the female rotor,

the distance between the farthest points of the adjacent liquid supply ports provided on the one side in the direction in which the screw rotor extends is smaller than the width of the tooth groove on the one side of the male rotor or the female rotor.

5. Liquid cooled screw compressor according to any of claims 1 to 3,

at least two of the plurality of liquid supply ports are provided on one side of the male rotor or the female rotor, and a distance between nearest points of adjacent liquid supply ports provided on the one side is larger than a width of the tooth space on the one side or a distance between farthest points of adjacent liquid supply ports provided on the one side is smaller than the width of the tooth space on the one side in a direction in which the screw rotor extends.

6. Liquid cooled screw compressor according to any of claims 1 to 3,

the rotor housing has a suction port at a position corresponding to an end of the screw rotor in a direction in which the screw rotor extends,

the position of the liquid supply port closest to the suction port among the plurality of liquid supply ports is separated from the suction port by a width of the tooth groove on the side of the male rotor or the female rotor where the liquid supply port closest to the suction port is provided.

7. Liquid cooled screw compressor according to any of claims 1 to 3,

the rotor case has a discharge port at a position corresponding to an end of the screw rotor in a direction in which the screw rotor extends,

the position of the liquid supply port closest to the discharge port among the plurality of liquid supply ports is not less than the width of the tooth groove on the side of the male rotor or the female rotor where the liquid supply port closest to the discharge port is provided.