JP3403452B2 - Oilless screw compressor - Google Patents

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION The present invention relates to screw compression,
In particular, the present invention relates to an oil-free screw compressor which is 37 kW or less and is used for a high pressure stage of a two-stage machine.

[0002]

2. Description of the Related Art An oil-free screw compressor having a rotor having a pair of twisted tooth spaces which mesh with each other and does not inject oil between them has hitherto been known as a minimum capacity machine.
Up to 5 kW has been put to practical use, and the structure of the compressor is described in JP-A-3-2900089.

Since no lubricating oil is injected between the rotors of the oilless screw compressor for the purpose of sealing, the efficiency of this type of compressor is as high as that of a tooth space (compression chamber) represented by a gap between the rotors. It is greatly affected by the leakage from the surrounding gap to the suction side or low pressure side tooth space. Generally, an oil-free screw compressor is used at a rotor peripheral speed of 50 m / s or more in order to overcome these leaks and obtain a certain efficiency.

[0004] As the capacity of the oil-free screw compressor becomes smaller, the peripheral speed cannot be reduced below the above-mentioned value in order to secure a certain efficiency, so a small rotor is used. . Along with this, the discharge capacity V _TH per one rotation of the rotor becomes smaller.

However, it is difficult to proportionally reduce the gap in the compressor due to the accuracy limit of the processing machine, and the ratio of the internal leakage amount to V _TH is relatively increased, and the efficiency is a medium capacity machine, or It will be lower than large capacity machines.

Speaking of an air compressor, one having an output of 15 to 22 kW class is a small capacity machine having a lower limit at present in view of practical efficiency. In this case, the medium capacity machine is 37
~ 55kW class, large capacity machine refers to 75kW or more.

Also in the two-stage compression type oil-free screw compressor, the high-pressure stage compressor has a V _TH of about 40% to 60% with respect to the low-pressure stage. As for the stage machine, the minimum capacity machine with the lower limit of 37 to 55 kW class is used.

[0008]

In order to improve the efficiency of the screw compressor, it is desirable to make the gap surrounding the tooth space forming the compression chamber as small as possible. But on the other hand,
If the gap is made too small, the temperature will rise during operation, the rotor will bend due to pressure, etc.
The phenomenon of contact between the rotor and the casing may occur and even lead to an inoperable state.

In order to overcome these problems, Japanese Patent Publication No. 61-47
As described in 992, means for correcting the thermal expansion due to the temperature rise of the rotor and designing the rotor tooth profile have been taken. However, there is a limit to the accuracy of machine tools that process rotors, casings, etc. that make up compressors at a practical cost, and secure a safe gap that includes variations in bearing clearance and variations in operating conditions. I need to put it.
This value is about the same as the gap to be secured for processing when processing a rotor having a diameter of about 160 mm and when processing a rotor having a diameter of about 60 mm. Therefore, the efficiency is small for a small-diameter rotor with a small V _TH . Inevitable decline.

On the other hand, when designing and manufacturing a small-capacity machine, the rotor specifications of the medium-capacity machine are proportionately reduced. In the case of a single-stage machine, the capacity of 7.5 kW to 11 kW class is 22 more.
The rotor diameter is 4 in the high-pressure stage of a 2-stage machine of ~ 37kW class.
It is about 0 to 50 mm, and the shaft diameter determined by the rotor tooth groove is 12 to 15 mm, which reduces the rigidity of the bearing and the deflection that occurs during assembly and operation due to the decrease in shaft rigidity. Securing a gap becomes even more difficult. Further, in the oil-free screw compressor, a long gas seal is required in the axial direction in order to prevent the lubricating oil from entering the compression chamber, which also causes a problem that the shaft becomes longer and the deflection becomes larger. is there.

Further, a conventional example will be described with reference to the drawings.

FIG. 12 shows the structure of a conventional oil-free screw compressor. The male rotor 1 and the female rotor 2 have bearings 7 a, 7 inside the casing 3 and the suction side casing 10.
It is rotatably supported by b, 8a and 8b. Further, the angular contact ball bearings 9a and 9b are attached to the discharge side shaft of the rotor, and
The position of the rotor in the thrust direction is fixed so as not to contact the inner wall of the rotor 10. Further, timing gears 5 and 6 are provided at the shaft ends of the male rotor 1 and the female rotor 2, and are assembled by adjusting the rotational phase so that the rotors do not come into contact with each other during operation. No oil is used between the rotors 1 and 2, and the compressor is heated to a high temperature by the heat generated by the compression. Therefore, the casing 2 is provided with a jacket 13 for flowing water or a cooling liquid to replace it. There is.

In order to prevent the oil that lubricates the bearings and gears from entering the compression chamber formed by the rotor and at the same time prevent the compressed gas from leaking to the outside, the gas seal 11a,
11b, 11c, 11d and oil removing devices 12a, 12
b, 12c and 12d are provided on the shafts of the respective rotors.

The male rotor is driven by the pinion gear 4 attached to the suction side shaft end, and sucks gas from the gas suction port 14 between the tooth spaces generated by the meshing with the female rotor.
After being compressed to a predetermined pressure, the compressed gas is discharged from the gas discharge port 15.

The prior art example of FIG. 12 shows a minimum output oilless screw compressor (15 kW for a single-stage machine, 45 kW class for a high-pressure stage of a two-stage machine) which is currently in practical use. The number of male rotor teeth Zm is 5, the number of female rotor teeth Zf is 6, the outer diameter Dm of the male rotor 1 is 63 mm, the length Lm of the male rotor 1 is 94.5 mm, and Lm / Dm = 1.5.
Is. The discharge capacity per revolution of this rotor is 125
ml / min. When used in a 15 kW class compressor, the male rotor rotational speed is about 18000 rpm, and the rotor peripheral speed Um is about 60 m / s.

As described above, in the case of the oil-free screw compressor, this level of Um is the lower limit peripheral speed at which significant reduction in efficiency can be prevented.

Here, in order to realize a compressor having a smaller capacity, for example, a single stage of 7.5 kW, or a high pressure stage of two stages of 22 kW class, a conventional example (15 kW) is proportionally set to 1. When the discharge capacity per revolution is scaled down to about half, the male rotor outer diameter is 50 mm and the male rotor length is 75 mm.
mm. Here, in order to prevent the efficiency from decreasing, the rotation speed of the male rotor is set to about 2 in order to set the peripheral speed to about 60 m / s.
It will be 3000 rpm. Furthermore, this compressor is a single stage 11
kW, in the high-pressure stage of the two-stage machine, when used at 30 kW or 37 kW, the rotation speed of the male rotor is 23000 x (11 /
In 7.5), ultra-high speed rotation of about 33700 rpm is required. As the rotor diameter decreases, the shaft diameter of the bearing also naturally decreases, and the shaft rigidity also decreases significantly.However, while keeping the gap between the male and female rotors to a minimum, the gas temperature due to compression (in a single stage 300 ° C or higher, 160 for 2-stage machine
30,000 rpm, taking into account the thermal expansion due to
To operate a compressor at a speed exceeding 10 mph requires very advanced design technology and manufacturing technology.

The above problem is, for example, only for 7.5 kW, 1
If a compressor is designed so that it is dedicated to 1 kW, it will be somewhat alleviated, but on the other hand, the reciprocating type is still the mainstream in this class of compressors. Even if it is assumed that the machine has been realized, it will not be practical in terms of cost due to extremely severe processing accuracy.

The present invention solves the above problems,
It is an object of the present invention to provide an oil-free screw compressor having a smaller capacity than the one currently put into practical use.

[0020]

In order to solve the above-mentioned problems, an oil-free screw compressor according to the present invention is characterized by what is stated in claims 1 to 4, and particularly, An oil-free screw compressor according to the invention of claim 1 as an independent claim is a oil-free screw compressor in which a male rotor and a female rotor are meshed with each other and a thrust bearing is arranged on the discharge side of the rotor. Type screw compressor is used for high pressure stage of two-stage machine with 37 kW or less, the ratio of rotor length to rotor outer diameter is set to about 1.0 or less, and it is generated by the meshing of male and female rotor teeth. Looking at the volume reduction curve of the tooth groove (compression chamber), the rotation angle from the start of volume reduction until the tooth groove is separated from the suction side end surface of the casing is the same as the angle of rotation after the separation. It is characterized in that it is smaller than twice the rotation angle until the bending. That is, specifically, the rotor diameter is set to about 60 mm to 100 mm in order to use the rotor shaft rigidity and the rotor tooth profile machining accuracy, which are currently in practical use, to reduce the VTH. In order to reduce the length of the rotor, the length of the rotor is not the currently used rotor diameter × about 1.2 to 1.5, but
Rotor diameter x 1.0 or less. In this way, the rotor length is shortened and the helix angle is made large so that the period during which the suction stroke takes the compression stroke does not occupy a large portion of the entire compression stroke, and the tooth space is sucked into the casing from the start of volume reduction. The rotation angle until it is separated from the side end surface is less than twice the rotation angle after separation until the tooth groove reaches the discharge port, and the above technology is applicable to rotor diameters of 100 mm or less. To do.

Furthermore, the shape of the starting position of the suction port is such that the suction volume of the tooth space increases from 0 and opens after a while. Further, the present compressor has a structure in which the heat generated by compression is dissipated to the atmosphere, and no jacket for cooling liquid is provided.

[0022]

By shortening the rotor length, it is possible to obtain a small V _TH without lowering the rotor peripheral speed and without lowering the shaft rigidity due to the reduction of the rotor shaft diameter. By increasing the twist angle, it is possible to shorten the time during which the tooth groove is in contact with the suction-side end face where a relatively large gap is usually provided, so that the efficiency can be improved.

Further, when the compressor is used on the high pressure side of the two-stage machine, the pressure in the tooth groove in which compression is started rises faster than that in the single-stage machine, so from this tooth groove to the preceding suction side. Leakage into the tooth space can be prevented by delaying the opening of the suction port.

Further, by constructing the highly efficient ultra-small oil-free screw compressor as described above and minimizing the heat generation, the compressor can be operated without cooling.

[0025]

Embodiments of the present invention will now be described with reference to the drawings. FIG. 1 shows a cross-sectional structure of a compressor according to an embodiment of the present invention, and FIG. 2 shows a sectional view taken along the line AA of FIG.

1 and 2, the male rotor 1 and the female rotor 2 are rotatably supported inside the casing 3 and the suction side casing 10 by bearings 7a, 7b, 8a, 8b. Further, combined angular ball bearings 9a, 9b are attached to the discharge side shaft of the rotor, and the position of the rotor in the thrust direction is fixed so that the rotor does not contact the inner walls of the casings 3, 10.

Timing gears 5 and 6 are provided at the shaft ends of the male rotor 1 and the female rotor 2 and are assembled by adjusting the rotational phase so that the rotors do not come into contact with each other during operation. .

In order to prevent the oil that lubricates the bearings and gears from entering the compression chamber formed by the rotor and at the same time prevent the compressed gas from leaking to the outside, the gas seal 11a,
11b, 11c, 11d and oil removing devices 12a, 12
b, 12c and 12d are provided on the shafts of the respective rotors.

The male rotor is driven by the pinion gear 4 attached to the suction side shaft end, and sucks gas from the gas suction port 14 between the tooth spaces formed by the meshing with the female rotor.
After being compressed to a predetermined pressure, the compressed gas is discharged from the gas discharge port 15.

The above construction is similar to that of the conventional example.

Thus, in one embodiment of the present invention, the same male rotor outer diameter (Dm = 63 mm) as the conventional machine is adopted,
Make the male rotor length Lm about half that of the conventional model, that is, Lm /
By setting Dm = 0.75, the rotor capacity can be reduced, and at the same time, the rotor peripheral speed of 60 m / s or more can be secured in the conventional level of the rotational speed range. That is, the ratio of the rotor length to the rotor outer diameter is about 1.0 or less.

However, sufficient efficiency cannot be obtained only by processing the rotor to half the length of the prior art example. Hereinafter, efficiency will be described with reference to the drawings.

In FIG. 3, the number of teeth of the male rotor tooth or lobe is 5
The number of teeth of the female rotor is 6, and the volume change of the tooth space caused by the meshing of the male and female rotors is calculated as the rotation angle X of the male rotor.
The horizontal axis represents m. In FIG. 3, Xm referred to here is 0 ° at the position of the tooth M of the male rotor 1 in FIG. 3-1 and the volume of the tooth space V is expressed with respect to Xm.

Further, FIG. 4 shows that the pressure rise of gas (air in this example) in the tooth space caused by the volume change is
Is shown on the horizontal axis.

This example is an example of a high-pressure stage of a two-stage compressor, and the suction pressure is 2.88 kgf / cm ² a discharge pressure 8.03.
It is kgf / cm ² a.

In the case of air, there is a relationship of π = v ^1.4 (1) between the compression ratio π and the volume ratio v.

Therefore, in the above example, π = 8.03 / 2.88 = 2.79 v = 2.79 ^{1 / 1.4} = 2.08 Further, assuming that the specific volume is V, V = 1 / v = 1 / 2.08 = 0.48 In FIG. 3, the discharge port is provided on the inner wall surface of the casing so that the tooth groove opens at the discharge port at a rotation angle corresponding to V = 48%.

Here, up to Xm = 72 degrees, the compression chamber (tooth groove) is a section in which the volume is reduced while having a portion in contact with the suction side casing 10.

In the case of the oilless screw compressor, the discharge gas temperature rises up to about 240 ° C. even in the case of the two-stage machine. The temperature causes the rotor to thermally expand, but it is necessary to set the end face gap between the rotor and the casing in consideration of this expansion amount. It is customary to set a large gap on the suction side end face because the distance from the thrust bearings 9a and 9b is longer than that on the discharge side end face, the thermal expansion amount is large, and the pressure increase is small. However, in realizing an ultra-small oilless screw compressor as in this example,
It is also necessary to consider the compression loss in the suction side end face gap.

When the helix angle β on the outer circumference of the male rotor is set to β = 55 ° as in the prior art, the pressure between the tooth spaces is about 5 at Xm = 72 ° as shown by the dotted line in FIG. .5
It has risen to kgf / cm ² a. The preceding tooth space has entered the suction stroke, and the pressure here is a suction pressure of 2.88k.
It is gf / cm ² a. Therefore, the pressure difference between the tooth space and the preceding tooth space is about 2.6 kgf / cm ^2, and the suction end face with a large gap must seal this pressure difference until this pressure difference is reached. Leakage to the side leading tooth space increases, which is the main factor that reduces the efficiency of the compressor. In other words, in the conventional technique, the compression is completed at Xm = 97 °, but the suction end face with much loss is involved in the compression stroke until Xm = 72 °.

In FIG. 3, the solid line is an example in which the twist angle of the outer circumference of the male rotor of the present invention is 69 °, and the Xm at which the compression is completed (the discharge pressure becomes 8.03 kgf / cm ² a) is Xm.
= 140 °. Also, Xm during the suction side end face involved in the sealing of the tooth space is also 72 °, and the section where the suction end face seals during the entire compression stroke can be reduced to about 50%. The pressure in the tooth space at Xm = 72 ° is 4.15k.
gf / cm only not increased to ² a, the pressure difference between the suction side tooth grooves preceding be 4.15 to 2.88 = 1.27 kg
It can be reduced to f / cm ² which is half that of the prior art, and improvement of efficiency can be achieved by reduction of internal leakage during compression. At the same time, the reduction of the leakage to the suction side tooth groove has a great effect on the improvement of the suction efficiency of the compressor.

In general, in order to reduce the leakage inside the compressor by this method, it is necessary to keep the section where the suction end face is related to the seal smaller than 2/3 at least during the entire compression process.

The above will be described with reference to FIGS. 10A to 10C and FIG.

FIG. 10A shows the change in the leak area of each gap portion of the compression chamber formed by the tooth groove for each rotation angle, with the male rotor rotation angle X _M as the horizontal axis. A _{S in the} figure
Leakage area of the portion that is sealed at the suction end face, A _C is leakage area between the rotor and the casing, A _R is the leakage area of the meshing portion between the male and female rotors, A _D is sealed with the discharge end face It is the leakage area of the part where it exists. This example is an example in which the number of male rotor teeth is 5, and when Xm = 72 °, the seal at the suction end face is eliminated. Further, in this example, the final pressure is reached and the pressure is completed at Xm = 140 °. Therefore, the section X _S = 72 ° in which the suction end face is involved in the seal of the tooth space, and the section X _R = 68 ° thereafter until reaching the discharge port.
Therefore, X _S / X _R = 1.06.

FIG. 10B shows the pressure increase in the tooth space with respect to Xm, which is the same as FIG. In the figure, the solid line is an example of β = 69 ° and the broken line is an example of β = 55 °.

FIG. 10C shows the instantaneous internal leakage amount at each rotation angle with respect to Xm. A (leakage area) in FIG. 10A at each Xm position and FIG. the product of the (B) P in (pressure) is represented as the leakage amount Q _L.

The total amount of leakage generated by one tooth groove during the compression process is represented by the area Q _LT of the hatched portion in FIG.

This is an example of the case of β = 55 ° (X _S / X _R
= 72/25 = 2.88), A in the section of X _S
In addition to the very large _S , most of the compression process is completed in this section, so the total leakage amount Q _{LT '} is shown in FIG.
It can be seen that the area of the broken line hatched area in (C) of FIG.

[0049] Figure 11, taken with X _S / X _R, the total leak rate Q _LT to the longitudinal axis on the horizontal axis, the change of the Q _LT is smaller in a section of the X _S / X _R = 0~2, When it becomes 2 or more, Q _LT increases greatly.

The smaller X _S / X _R is, the better it is. However, it cannot be extremely reduced structurally, and if it is 2 or less, the performance is not deteriorated.

A method for reducing leakage at the suction end face for the same purpose will be described with reference to FIGS.

FIG. 5 shows a prior art suction port configuration S, where the tooth position of the 1M male rotor is called Xm = 0 °, ie, one lobe or tooth of the male rotor as seen from the end. The rotation angle of the male rotor at the time when the outer shape of the end substantially coincides with the outer shape of the end of one tooth groove of the female rotor is 0 °. This 1M
The tooth space P adjacent to is the position where the compression starts, and the suction port S is closed at this point.

On the other hand, the starting shape Q of the suction port S is Xm = 0.
The tooth groove opens to the suction port S at the same time that the suction tooth groove volume appears, assuming that the rotor has an advancing surface lobe shape at °.

On the other hand, in FIG. 6, Xm is 15 ° from FIG.
Although the meshing at a position delayed by a certain degree is shown, the portion where the tooth groove in which compression has already started contacts the suction side end face is the crescent-shaped portion H in the figure.

FIG. 7 shows a detailed explanatory view of this portion. The portion H which is compressed and is in contact with the suction end face moves upward in the drawing as the area thereof decreases as the rotor rotates to show a plurality of crescent shapes in the drawing. The portion H _{E where} the pressure rises most and the start line Q of the suction port are almost at the same position. Usually, if the opening line Q formed in the casting shape is slightly displaced, the pressure difference is about 1 to 2 kgf as described above. The air in the tooth space where the compression has already started leaks to the preceding suction side tooth space with / m ² g.

Therefore, in the present invention, as shown in FIG. 8, the shape K of the rotor advancing surface at the rotational position where the male rotor 1 advances 5 ° from the position of Xm = 0 on the opening line Q of the intake port.
Form along the shape. Or a distance equivalent to this H
Form with a curve or straight line away from _E. As a result, leakage from H _E can be greatly reduced.

Further, as described above, a casing 3 including a rotor, which constitutes a non-conventional small-sized and small-capacity oilless screw compressor, reduces heat generation due to internal leakage between rotors, and includes the rotors, is shown in the figure. The structure does not have a jacket for flowing the cooling liquid, and it is naturally cooled or air cooled. In the present invention, since the shaft diameter equivalent to that of the conventional minimum output machine can be maintained, the outer surface area of the casing is not reduced in proportion to the reduction in discharge capacity.

Further, due to the above-described efficiency improvement, the heat generated by the internal leakage of the compressor is small, so that the heat radiation by the cooling liquid is not necessary, and the simplification of the structure is largely achieved.

[0059]

According to the present invention, ultra-small, the oil-free screw compressor of the high pressure stage follows 22~37kW class as high pressure stage immediate Chi 2-stage machine, reduction in efficiency in the state of the machining precision level It can be provided and realized without imitating reliability and without lowering reliability.

[Brief description of drawings]

FIG. 1 is a cross-sectional view of one embodiment of an oil-free screw compressor of the present invention.

FIG. 2 is a sectional view taken along line AA of FIG.

FIG. 3 is an explanatory diagram of volume change curves of the present invention and a conventional example.

FIG. 4 is an explanatory diagram of acupressure lines according to the present invention and a conventional example.

FIG. 5 is an explanatory diagram of a conventional suction port.

FIG. 6 is a diagram similar to FIG. 5, showing a rotor position delayed by 15 degrees.

FIG. 7 is a detailed view of FIGS. 5 and 6;

FIG. 8 is an explanatory view of an intake port of the present invention.

FIG. 9 is a diagram showing the relationship between the male rotor rotation angle and the change in the leak area of each gap portion.

FIG. 10 is a diagram showing a male rotor rotation angle, a leakage area, a pressure increase in a tooth space and a leakage amount.

FIG. 11 is a diagram showing a relationship between X _S (section where suction end surface is related to seal of tooth groove) / X _R (section where suction end surface reaches discharge port) and total leakage amount.

FIG. 12 is a horizontal sectional structural view of a conventional screw compressor.

FIG. 13 is a vertical sectional structural view of a conventional screw compressor.

[Explanation of symbols]

1 ... Male rotor 2 ... Female rotor 3 ... Casing 4 ... Pinion gears 5, 6 ... Timing gears 7a, 7b, 8a, 8b ... Radial bearings 9a, 9b ... Thrust bearing 14 ... Suction port 15 ... Discharge port

─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Toshikazu Uchida 502 Jinrachi-cho, Tsuchiura-shi, Ibaraki Machinery Research Laboratory, Hitachi, Ltd. (56) References JP-A-4-159488 (JP, A) JP-A-4 −209980 (JP, A) (58) Fields investigated (Int.Cl. ⁷ , DB name) F04C 18/16

Claims (4)

(57) [Claims] 1. A non-lubricating screw compressor in which a male rotor and a female rotor are meshed with each other and a thrust bearing is disposed on the discharge side of the rotor, wherein the non-lubricating screw compressor is 37 kW or less and is a two-stage machine. It is used for the high pressure stage, the ratio of rotor length to rotor outer diameter is set to about 1.0 or less, and the volume decrease curve of the tooth groove (compression chamber) caused by the meshing of male and female rotor teeth The rotation angle from the start of the decrease until the tooth groove is separated from the suction-side end surface of the casing is less than twice the rotation angle from the separation until the tooth groove reaches the discharge port. Refueling screw compressor. 2. The oilless screw compressor according to claim 1, wherein the outer diameter of the rotor is 100 mm or less. 3. The tooth opening (compression chamber) has a volume based on a rotation angle of 0, and the opening position of the suction port corresponds to a rotation angle of the rotor of 5 degrees.
The oil-free screw compressor according to claim 1 or 2, wherein the position is advanced by more than one degree. 4. The oilless screw compressor according to claim 1, wherein the casing including the rotor does not have liquid cooling means.