Screw compressor

A screw compressor comprises a casing and a pair of male and female rotors respectively having herical teeth and grooves formed in their outer peripheries. The rotors are arranged within the casing in meshing relation to each other through the teeth and grooves. Each rotor is rotatably supported by a pair of first and second radial plain bearings and a thrust plain bearing which is arranged at an end of the rotor shaft and adjacent the second radial plain bearing. Each of the thrust plain bearings is of dynamic-pressure type in which with rotation of a corresponding one of the shafts of the respective rotors, pressure is applied to lubricant by sliding movement of the thrust plain bearing to produce an anti-thrust force. The lubricant for each of the second radial plain bearings and for a corresponding one of the thrust plain bearings is fed to a space between the second radial plain bearing and the thrust plain bearing, and is distributed to the second radial plain bearing more in amount than that fed to the thrust bearing due to a difference between flow resistance with which the lubricant passes through the second radial plain bearing and flow resistance with which the lubricant passes through the thrust plain bearing.

BACKGROUND OF THE INVENTION 
The present invention relates to an improvement in bearings for use in a 
screw compressor. 
In general, the arrangement of a screw compressor is such that a pair of 
male and female rotors rotate in meshing relation to each other through 
teeth and grooves to compress gas such as refrigerant gas or the like. 
Because of such arrangement, large force acts upon the rotors, and the 
rotors are required to rotate at a high speed. The screw compressor is 
provided with radial bearings for supporting radial force from the rotors 
and thrust bearings for supporting axial force from the rotors. 
Conventionally, angular-contact rolling bearings have been employed for the 
purposes described above, and have been arranged in plural to increase the 
load capacity. As a result, the bearings occupy much space so that the 
entire compressor tends to be brought to a large size. In addition, in a 
hermetic screw compressor, a semipermanent service life is required for 
each bearing. The rolling bearing is advantageous in that it is hard to 
seize up even under inferior lubricating conditions. However, the rolling 
bearing have a relatively short service life because of rolling fatigue. 
Screw compressors employing rolling bearings are described, for example, in 
Japanese Patent Application Laid-Open Nos. 57-119191 and 57-206791. 
In order to solve the problems discussed above, it is desirable that slide 
bearings or plain bearings are employed as a substitute for the screw 
compressor in rolling bearings. The plain bearing is advantageous in that 
the structure is compact, the cost is low, the service life can be 
prolonged, and the like. On the other hand, the plain bearing is subject 
to seizing and becomes unable to rotate, if formation of a lubricating oil 
film is insufficient. 
Gas within the screw compressor is brought to a high temperature level in 
the course of compression of the gas, so that the lubricating oil often 
reaches a temperature level within a range of from 130 to 140 degrees C. 
Further, the lubricating oil is mixed with the compressed gas such as 
refrigerant or the like and tends to be reduced in viscosity, and the load 
applied to the bearings is high. In particular, if plain bearings are 
employed in a low-capacity hermetic screw compressor which utilizes a 
differential-pressure oil supply system, some of the plain bearings would 
seize up in a moment unless a sufficient amount of lubricating oil is fed 
to all of the plurality of bearings. 
Japanese Patent Application Laid-Open No. 56-9694 discusses problems 
concerning the plain bearing lubrication in a screw compressor employing 
the differential-pressure oil supply system, and has proposed one solution 
to of the problems. According to this prior art, the lubricating oil is 
fed to a substantially central portion of each of two radial bearings for 
each of two rotor shafts. At the same time, in order to secure lubrication 
of each thrust bearing for the rotor shaft, the lubricating oil is 
supplied to a hydraulic pressure chamber defined between each thrust 
bearing and the adjacent radial bearing, through a passageway having 
provided therein a restriction, and the lubricating oil is fed from the 
hydraulic pressure chamber to the thrust bearing. 
The solution proposed in the above prior art can get considerable results. 
If consideration is made to the inferior lubricating conditions as 
described above, however, it is desirable to further enhance the 
lubricating ability, in particular, of the thrust bearings. A plain 
bearing of dynamic-pressure type in which pressure is applied to the 
lubricating oil at a sliding section is superior in the ability of forming 
an oil film, and is effective for this purpose. However, supply of the 
lubricating oil according to the aforesaid prior art is inadequate for the 
case of employment of dynamic-pressure type thrust bearings, since supply 
of the lubricating oil to each thrust bearing and supply of the 
lubricating oil to the radial bearings are brought to an extremely 
unbalanced state. 
That is, in case where the dynamic-pressure type thrust bearings are 
employed in a screw compressor and the lubricating-oil supply means as 
disclosed in Japanese Patent Application Laid-Open No. 56-9694 an amount 
of the lubricating oil passing through each thrust plain bearing is low, 
because the resistance to flow of the lubricating oil through the thrust 
plain bearing is high. On the other hand, the path guiding the lubricating 
oil to the central portion of the radial plain bearing adjacent to each 
thrust plain bearing is low in flow resistance. Accordingly, the 
lubricating oil almost does not flow to a section of the radial plain 
bearing on the side of the thrust bearing, but only a section of the 
radial plain bearing on the side opposite to the thrust bearing is 
lubricated. For this reason, the section of the radial plain bearing on 
the side of the thrust bearing seizes in a short period of time. 
SUMMARY OF THE INVENTION 
It is an object of the invention to provide a screw compressor in which the 
seizure-resistant ability of plain bearings is improved, the structure is 
compact, and the service life of the bearings is long. 
To this end, the invention employs dynamic-pressure type plain bearings as 
thrust bearings for the screw compressor, and conversely utilizes an 
imbalance in flow resistance between the dynamic-pressure type plain 
bearings and radial plain bearings to thereby improve the lubricating 
ability of the entire plain bearings. 
According to the invention, there is provided a screw compressor comprising 
a casing, a pair of male and female rotors respectively formed in outer 
peripheries thereof with helical teeth and grooves, with the male and 
female rotors being arranged within the casing in meshing relation to each 
other through the teeth and grooves, two pairs of first and second radial 
plain bearings for respectively supporting shafts of the male and female 
rotors, and a pair of thrust plain bearings arranged respectively adjacent 
the second radial plain bearings for supporting one ends of the shafts of 
the male and female rotors. Each of the thrust plain bearings is of 
dynamic-pressure type in which with rotation of a corresponding one of the 
shafts of the male and female rotors, pressure is applied to lubricant at 
a sliding section of the thrust plain bearing to produce an antithrust 
force, and the lubricant is fed to each of the first radial plain 
bearings, and is fed to a space between an end of a sliding section of 
each of the second radial plain bearings and an adjacent end of the 
sliding section of a corresponding one of the thrust plain bearings so 
that the lubricant is distributed to the second radial plain bearing and 
the thrust bearing to lubricate these bearings in proportion to a 
difference between flow resistance with which the lubricant passes through 
the second radial plain bearing and flow resistance with which the 
lubricant passes through the thrust plain bearing. 
In the above arrangement, the lubricant fed to the space between each 
second radial plain bearing and a corresponding one of the thrust plain 
bearings flows in part to the thrust plain bearing, and the remaining 
lubricant flows from one end to the other end of the second radial plain 
bearing to lubricate its entire axial length. A requisite amount of 
lubricant for lubricating each thrust plain bearing is determined 
depending upon rotational speed of a corresponding one of the rotors and 
compressing conditions. Since, however, each thrust plain bearing is of 
dynamic-pressure type, a thick lubricant film is formed with a less amount 
of the lubricant, at the sliding section of the thrust plain bearing, 
making it possible to prevent seizing-up of the thrust plain bearing. 
On the other hand, since each radial plain bearing is lubricated by the 
remaining rich amount of lubricant, the radial plain bearing is prevented 
from seizing up even under a wide range of operating conditions of the 
screw compressor so that the radial plain bearing operates in a stable 
fashion.

DETAILED DESCRIPTION OF THE EMBODIMENTS 
The invention will be described below in detail with reference to 
embodiments illustrated in the accompanying drawings. 
Referring to FIG. 1, a screw compressor according to an embodiment of the 
invention comprises a tubular rotor casing 2 and an electric motor 1 
mounted to an upper end of the rotor casing 2. A cup-shaped suction casing 
3 is gas-tightly mounted to the rotor casing 2 by bolts or the like to 
cover the electric motor 1. 
A pair of male and female rotors 4 and 5 are rotatably accommodated in the 
rotor casing 2. The male and female rotors 4 and 5 are respectively formed 
in their outer peripheral surfaces with helical teeth and grooves. The 
rotors 4 and 5 are arranged in meshing relation to each other through the 
teeth and grooves. A suction inlet 6 is provided in the suction casing 3. 
Gas, compressed as refrigerant or the like passing through the suction 
inlet 6, is introduced into the mating section between the male and female 
rotors 4 and 5 through the electric motor 1. A discharge outlet 7 is 
provided in the rotor casing 2 in communication with the mating section 
between the male and female rotors 4 and 5 to permit the compressed gas to 
be discharged through the discharge outlet 7. 
The male and female rotors 4 and 5 have their respective one shafts 4a and 
5a which are supported respectively by radial plain bearings 8a and 8b 
arranged in the rotor casing 2. The shaft 4a is connected to the electric 
motor 1. The male and female rotors 4 and 5 have their respective other 
shafts 4b and 5b which are supported respectively by radial plain bearings 
9a and 9b mounted also to the rotor casing 2. A pair of thrust plain 
bearings are arranged respectively at ends of the respective shafts 4b and 
5b for receiving axial forces acting upon the respective male and female 
rotors 4 and 5. 
As shown in FIG. 2 in an enlarged fashion, each of the thrust plain 
bearings is composed of a disc-like runner 11a, 11b fixedly mounted to a 
corresponding one of the shafts 4b and 5b of the respective male and 
female rotors 4 and 5, and a disc-like stationary bearing element 10a, 10b 
arranged in facing relation to the runner 11a, 11b. Each of the stationary 
bearing elements 10a and 10b is formed therein with a through bore. The 
stationary bearing elements 10a and 10b are fixedly mounted to the rotor 
casing 2 and are arranged respectively above the runners 11a and 11b in 
such a manner that the shafts 4b and 5b of the respective rotors extend 
respectively through the through bores in the respective stationary 
bearing elements 10a and 10b. As the screw compressor operates, the male 
and female rotors 4 and 5 are moved upwardly under the reaction from 
compression of the gas to cause the runners 11a and 11b to be urged 
respectively against the stationary bearing elements 10a and 10b. 
A planar element 12a and a ball element 13a are arranged at the shaft end 
of the male rotor 4 in a close relation to the thrust plain bearing 
therefor. Likewise, a planar element 12b and a ball element 13b are 
arranged at the shaft end of the female rotor 5 in a close relation to the 
thrust plain bearing therefor. The planar element 12a is in the form of a 
cup and is threadedly engaged with the shaft end of the male rotor 4. The 
planar element 12b is formed by one end of a positioning bolt 14b which is 
mounted to a bottom of a cover plate 16 by a lock nut 15b, with the planar 
element 12b facing toward the shaft end of the female rotor 5. The cover 
plate 16 is fixedly mounted in a gas-tight fashion to a lower face of the 
rotor casing 2 so as to cover the shaft ends of the respective male and 
female rotors 4 and 5. Threadedly engaged with the shaft end of the female 
rotor 5 is a runner retainer 11c in the form of a bolt which is arranged 
in facing relation to the positioning bolt 14b. The runner retainer 11c 
has a lower end face formed with a recess. The ball element 13b is 
received in the recess and is in contact with the planar element 12b. 
Also at a position below the planar element 12a on the side of the male 
rotor 4, a positioning bolt 14a is mounted to the bottom of the cover 
plate 16 by a lock nut 15a in facing relation to the planar element 12a. 
The ball element 13a is received in a recess formed in a head of the 
positioning bolt 14a, and is in contact with a lower surface of the planar 
element 12a. The planar elements 12a and 12b and the ball elements 13a and 
13b rotatably support the male and female rotors 4 and 5 when the reaction 
force acting upon each of the rotors due to the compressed gas is lower 
than the weight of the rotor, at start-up or at stopping of the screw 
compressor. 
A cup-shaped high-pressure gas casing 19 is gas-tightly mounted to the 
lower portion of the rotor casing 2 by means of bolts or the like. The 
high-pressure casing 19 cooperates with the suction casing 3 to form a 
pressure vessel. The high-pressure casing 19 has an interior thereof 
serving as an oil reservoir 21 in which the lubricating oil 20 is stored. 
On the other hand, a pair of strainers 18a and 18b are attached to the 
lower end of the cover casing 16, and have their respective lower ends 
immersed in the lubricating oil 20. 
A pair of feed-oil bores 17a and 17b are formed which extend through the 
respective strainers 18a and 18b, the cover casing 16 and the rotor casing 
2. The feed-oil bore 17a connects the oil reservoir 21 to the radial plain 
bearings 8a and 8b on the suction side, while the feed-oil bore 17b 
connects the oil reservoir 21 to a pair of spaces or hydraulic pressure 
chambers 17c and 17d (see FIG. 2) between the respective radial plain 
bearings 9a and 9b on the discharge side and the respective thrust plain 
bearings. 
In operation of the screw compressor according to the illustrated 
embodiment, as the electric motor 1 is turned on, the rotors 4 and 5 are 
rotatively driven by the electric motor 1. As indicated by the arrows in 
FIG. 1, the gas flowing into the suction casing 3 through the suction 
inlet 6 passes through the electric motor 1 while cooling the same. The 
gas is drawn into the mating section between the rotors 4 and 5 through a 
suction port (not shown) provided in the rotor casing 2 and is compressed 
by the rotors 4 and 5. The compressed gas is discharged to the interior of 
the high-pressure casing 19 through a discharge port (not shown) provided 
in the rotor casing 2. Subsequently, the compressed gas passes through an 
oil separator (not shown) and flows out of the compressor through the 
discharge outlet 7. 
As the screw compressor operates, high pressure is applied to the 
lubricating oil 20 within the oil reservoir 21. On the other hand, the 
opposite ends of each of the radial plain bearings 8a and 8b on the 
suction side, the upper ends of the respective radial plain bearings 9a 
and 9b, and a space within the cover casing 16 serving as a drain oil 
chamber 16a communicate with the suction side of the rotors and are 
maintained at a pressure lower than that within the high-pressure casing 
19. The lubricating oil 21 within the oil reservoir 20 is delivered to 
these bearings by a differential pressure between the discharge pressure 
and the suction pressure of the gas. The lubricating oil is fed to the 
radial plain bearings 8a and 8b on the suction side through the feed-oil 
bore 17a, and to the hydraulic pressure chambers 17c and 17d through the 
feed-oil bore 17b. The lubricating oil fed to each of the hydraulic 
pressure chambers 17c and 17d flows in part to a corresponding one of the 
thrust plain bearings, while the remaining lubricating oil flows to a 
corresponding one of the radial plain bearings 9a and 9b on the discharge 
side. The lubricating oil flowing to each thrust plain bearing lubricates 
a bearing face of a corresponding one of the stationary bearing elements 
10a and 10b and a bearing face of a corresponding one of the runners 11a 
and 11b. Subsequently, the lubricating oil is discharged into the drain 
oil chamber 16a. 
The thrust plain bearings employed in the invention are of dynamic pressure 
type. In the illustrated embodiment, each of the runners of the respective 
thrust plain bearings is formed with spiral grooves. Specifically, as 
shown in FIG. 3, each runner 11a, 11b has an end face in contact with a 
corresponding one of the stationary bearing elements 10a and 10b, which 
end face is provided with a plurality of spiral grooves 11x each having a 
predetermined depth, with a land 11y left between each pair of adjacent 
grooves. The spiral grooves 11x extend from a central portion of the 
runner to the periphery thereof along the rotational direction of the 
runner indicated by the arrow in FIG. 3. Further, each runner 11a, 11b is 
formed at its central portion with an annular oil groove 11z. On the other 
hand, each stationary bearing element 10a, 10b has the bearing face which 
is formed smoothly. 
During operation of the screw compressor, the lubricating oil fed through 
the feed-oil bore 17b and the hydraulic pressure chambers 17c and 17d 
reaches the bearing faces of the respective thrust plain bearings through 
annular gaps defined respectively between the through bores in the 
respective stationary bearing elements and the rotor shafts. The 
lubricating oil tending to flow out of the bearing face of each of the 
thrust plain bearings is confined in the spiral grooves 11x of a 
corresponding one of the runners 11a and 11b with rotation of the latter. 
As a result, dynamic pressure is generated in the lubricating oil within 
each thrust plain bearing due to the spiral grooves 11x of a corresponding 
one of the runners, so that the runner is floated away from a 
corresponding one of the stationary bearing elements by the oil film of 
high pressure against the upward axial force acting upon a corresponding 
one of the rotors. In this manner, the lubricating oil having lubricated 
the thrust plain bearings is discharged into the drain oil chamber 16b, 
but the discharged oil is extremely small in amount. This is because the 
dynamic pressure generated by the spiral grooves 11x serves as a 
resistance to hinder flow of the lubricating oil. For this reason, a high 
load capacity is produced in each bearing. 
Referring to FIG. 2, since the high flow resistance is caused at the thrust 
plain bearings as described above, almost all of the lubricating oil 
supplied to the hydraulic pressure chambers 17c and 17d flow toward the 
upper radial plain bearings 9a and 9b to lubricate their respective 
bearing faces. Subsequently, the lubricating oil discharged through the 
upper ends of the respective radial plain bearings 9a and 9b. The 
lubricating oil passes through a guide bore (not shown) and is finally 
returned to the oil reservoir 21 through the space on the suction side. In 
this manner, the lubricating oil having lubricated the radial plain 
bearings 9a and 9b is recirculated without staying. 
As shown in FIGS. 4 and 5, each of the radial plain bearings 9a and 9b on 
the discharge side is provided in its bearing face with an axial groove. 
The axial groove is composed of a relatively wide groove section 9x and a 
relatively narrow groove section 9y which is located on the rotor side. 
The narrow groove section 9y serves as a resistance to flow of the 
lubricating oil lubricating the radial plain bearing 9a, 9b, so that the 
lubricating oil passing through the narrow oil groove 9y experience a 
pressure drop. Although the amount of oil lubricating the radial plain 
bearing 9a, 9b is rich as described above, provision of the narrow oil 
groove 9y shown in FIG. 4 makes it possible to improve the load capacity. 
Each of the radial plain bearings 8a and 8b on the suction side is also 
formed in its bearing face with an axial groove, as shown in FIGS. 6 and 
7. Since the pressure around the radial plain bearings 8a and 8b on the 
suction side is the low suction pressure and has a uniform pressure 
distribution, a superior lubricating condition is obtained if the oil is 
fed to an axially central portion of the bearing face of each of the 
radial plain bearings. Further, the oil having lubricated the radial plain 
bearings 8a and 8b is easy to flow out because the surrounding pressure is 
low, so that there is a tendency that the lubricating oil fed is excessive 
in amount. Accordingly, the axial groove of each radial plain bearing is 
composed of a pair of relatively narrow groove sections 8y and 8y provided 
respectively at the opposite axial ends of the bearing face and a 
relatively wide groove section 8x connecting the narrow groove sections to 
each other. A feed-oil bore 8z is provided in the vicinity of the axial 
center of the groove section 8x. Each narrow groove section 8y has a 
cross-sectional configuration which is narrower than the narrow groove 
section 9y formed in the bearing face of each radial plain bearing on the 
discharge side, in order to prevent the lubricating oil fed to the radial 
plain bearings on the suction side from becoming excessive in amount. This 
is because the radial plain bearings 8a and 8b on the suction side are 
lower in bearing load than the radial plain bearings 9a and 9b on the 
discharge side so that the radial plain bearings 8a and 8b on the suction 
side can be lubricated with a less amount of oil. Moreover, the reason for 
the above is that if an excessive amount of lubricating oil flow to the 
radial plain bearings 8a and 8b, the lubricating oil fed to the radial 
plain bearings 9a and 9b on the discharge side is reduced in amount so 
that there is an anxiety causing seizing-up of the latter plain bearings 
which require the high load capacity. 
The differential-pressure oil supply system is high in reliability and 
compact in size, because an external oil supply source such as a pump or 
the like is dispensed with. However, the differential-pressure oil supply 
system is restricted by an oil supply amount. According to the invention, 
the thrust plain bearings, which are most severe in operating condition of 
the screw compressor, can be lubricated with a small feed amount of oil, 
so that it is possible for various bearings different in rotational 
condition from each other, to be efficiently lubricated with the 
restricted total feed amount of oil in the differential-pressure oil 
supply system. 
It is to be understood that the invention is not limited in application 
only to the screw compressor of differential-pressure oil supply system. 
That is, the invention can secure superior lubrication of bearings due to 
adequate distribution of lubricating oil, even in a compressor of powered 
oil supply system by a pump or the like. It will also be understood that 
other various modifications can be made to the invention within the scope 
of the appended claims. 
FIG. 8 shows a principal portion of a screw compressor according to another 
embodiment of the invention. In the embodiment, a radial plain bearing 91 
on the discharge side is formed with a cut-out 22 at an end of a bearing 
face of the bearing 91 on the side of the hydraulic pressure chamber 17d. 
Further, the feed-oil bore 17e communicates with the cut-out 22 so that 
the lubricating oil is fed from the cut-off 22. Other construction and 
operation of the embodiment illustrated in FIG. 8 are similar to those of 
the afore-mentioned embodiment. Like or similar component parts are 
designated by the same reference numerals, and the description of such 
similar component parts will therefore be omitted. According to the 
embodiment illustrated in FIG. 8, it is easy to process the feed-oil bore, 
making it possible to reduce the flow passage resistance of the feed-oil 
bore. 
If the radial plain bearings on the discharge side are replaced by bearings 
of a dynamic pressure type, the load capacity of the bearings increases 
like the thrust plain bearings of the invention. This makes it possible to 
reduce a consumptive amount of the lubricating oil. 
As described above, according to the invention, the lubricant is divided 
into two flows which respectively lubricate the radial plain bearings on 
the discharge side and the thrust plain bearings. Since the thrust plain 
bearings are of dynamic pressure type, it is possible to prevent seizing 
of the thrust plain bearings even if an amount of lubricating oil fed to 
the bearings is small. Further, since the radial plain bearings are 
lubricated by the remaining rich amount of oil, the radial plain bearings 
are prevented from seizing under the wide range of operating conditions of 
the screw compressor, making it possible to obtain its stable function or 
operation. Thus, there are obtained such advantages that the entire 
bearings in the screw compressor can be made compact in size, and can have 
their respective semipermanent service lives extended.