JP2981276B2 - Multi-stage screw compressor - Google Patents

Description

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a multi-stage screw compressor in which a fluid compressed by a low-pressure compressor is sucked into a high-pressure compressor and compressed to a predetermined pressure. The present invention relates to a multi-stage screw compressor suitable for reducing fluctuations in a discharge pressure of a side compressor and a suction pressure of a high-pressure side compressor.

[Prior Art] A conventional technique will be described with reference to FIGS.

FIG. 13 is a longitudinal sectional view of a two-stage screw compressor. In FIG. 13, the first stage compressor 1 which is a low pressure side compressor is shown.
And the two-stage compressor 2, which is a high-pressure compressor, are each composed of a pair of male rotors and female rotors, and these rotors are provided in bearings 6 to 6 provided in casings of the respective compressors.
9, the male and female rotors rotate while meshing with each other.

The male rotor 3 of the first-stage compressor 1 and the male rotor 4 of the second-stage compressor 2 are connected at their respective shaft ends by a gear coupling 11, and the motor 14 drives the first-stage compressor 1. By doing so, each compressor rotates.

In this two-stage screw compressor, the fluid sucked from the suction port 13 of the first-stage compressor 1 is compressed to an intermediate pressure, discharged from the discharge port 12 of the first-stage compressor 1, and passes through the intermediate casing 5. To the second stage compressor 2. Here, it is compressed and discharged to a predetermined discharge pressure.

FIG. 14 is a transverse sectional view (a sectional view taken along line XIV-XIV in FIG. 13) of the discharge port portion of the first-stage compressor 1. FIG. 15 is a cross-sectional view (XV-XV in FIG. 13) of the two-stage compressor 2 on the suction port side.
FIG.

FIG. 16 to P _'1 along the outer periphery from the intersection P ₁ of the male rotor 3 and the outer periphery of the outer periphery and the female rotor 15 of the male rotor 3 in one stage side compressor 1 shown in FIG. 14, also the intersection point P ₁ From female rotor
FIG. ₁₃ is a helix development view showing a plane up to P ′ ₁ along the outer circumference of _No. 15; Several lines shown diagonally in FIG. 16 show lines along the torsional direction of the tooth tips of the male rotor 3 and the female rotor 15, and with the rotation of the rotor, in the direction shown by the arrow in FIG. The line moves from the suction side to the discharge side. The end points m ₁ and f ₁ of the discharge port in the axial direction in FIG. 14 correspond to the points m ₁ and f ₁ in FIG. 16, respectively.

The tooth tip lines of the male rotor 3 and the female rotor 15 are respectively
The time when m ₁ and f ₁ are reached is the moment when the discharge port 12 is opened, and the fluid compressed between the rotor tooth spaces is discharged from the discharge port 12.

Figure 17 is up to P _'2 along the outer periphery from the intersection P ₂ of the male rotor 4 and the outer circumference of the outer periphery and the female rotor 16 of the male rotor 4 in the two-stage compressor 2 shown in FIG. 15, also the intersection point P ₂ From female rotor
FIG. 16 is a helix development view showing a plane up to P ′ ₂ along the outer circumference of 16; Similar to FIG. 16, the line of the tooth tip of the rotor moves from the suction side to the discharge side in the direction indicated by the arrow in FIG. 17 with the rotation of the rotor. In FIG. 17, m ₂ and f ₂ indicate the end point of the suction port. When the tip of the rotor reaches this point, the suction port is closed, that is, the suction is completed. The suction air as the fluid to be compressed is drawn into the tooth space between the male rotor 4 and the female rotor 16 through the suction port 18.

FIG. 18 is a volume curve showing the degree to which the volume between one tooth groove of engagement between the male rotor and the female rotor changes with respect to the rotation angle of the male rotor on the suction side and the discharge side. As shown in FIG. 18, the volume changes linearly with respect to the rotation angle of the male rotor in the range from −300 ° to −60 ° on the suction side and from 60 ° to 300 ° on the discharge side. . In other rotation ranges, it does not change linearly.

FIG. 19 is a diagram showing the change of the suction volume with respect to the rotation angle of the male rotor of the two-stage compressor in the case of the rotor showing the volume curve of FIG. Here, since the inter-tooth space volume of three teeth of the male rotor is synthesized at the suction port, the suction volume V on the vertical axis in FIG. This is a value synthesized by shifting each other. As can be seen from FIG. 19, the suction volume does not have a linear characteristic with respect to the rotation angle θ of the male rotor.

FIG. 20 is a diagram showing how the value of the change amount dv of the suction volume (= dv / dθ) with respect to the change dv / dθ of the rotation angle of the male rotor in FIG. 19 changes depending on the rotation angle of the male rotor. It is. This dv / dθ is called the suction volume gradient.
This corresponds to the amount of air sucked in per unit time because the rotation angle of the male rotor is constant. According to FIG. 20, when the rotation angle of the male rotor is in the range of −120 ° to −60 °, the suction volume gradient increases as the rotation proceeds. In other words, the amount of air taken in per unit time increases. When the rotation angle of the male rotor is in the range of -60 ° to 0 °, the suction volume gradient decreases as the rotation proceeds. In other words, the amount of air taken in per unit time falls as the rotation progresses. FIG. 19 and FIG.
The rotation angle of the male rotor in FIG. 20 is 0 ° when the suction is completed.

FIG. 21 is a diagram showing a change in the discharge volume with respect to the rotation angle of the male rotor of the one-stage compressor in the case of the rotor having the characteristic curve of FIG. Assuming that the start of discharge is 120 ° as a reference for the rotation angle of the male rotor from FIG. 21, the discharge volume changes linearly when the rotation angle of the male rotor is in the range of 0 to 90 °, but the rotation volume is 90 to 120 °. It does not change linearly in the range.

FIG. 22 is a diagram showing how the value (= dv / dθ) of the change amount dv of the discharge volume with respect to the change amount dθ of the rotation angle of the male rotor shown in FIG. 21 changes according to the rotation angle of the male rotor. is there. This dv / dθ is referred to as a discharge volume gradient, which corresponds to the amount of air discharged per unit time. According to FIG. 22, assuming that the discharge start time is 120 ° as a reference of the rotation angle of the male rotor, the rotation angle of the male rotor is 0 to 9
The volume gradient is constant in the range of 0 °, but 90 to 120
It can be seen that in the range of ゜ (the range of 30 ° before the opening of the discharge port to start the discharge), it is not constant but drops.

The intermediate pressure is constant when the suction volume gradient of the second stage compressor in FIG. 20 and the discharge volume gradient of the first stage compressor in FIG. 22 are always equal. The difference causes the intermediate pressure to fluctuate.

By the way, in the prior art, the rotation angle of the male rotor 3 of the first-stage compressor 1 which is the low-pressure compressor and the rotation angle of the high-pressure compressor 2
The stage-side compressor 2 does not have a structure in which the phase difference with the rotation angle of the male rotor 4 is fixed at a predetermined position. For this reason, the structure is such that any phase difference can be attached depending on the attachment method. As a result, for example, the first stage compressor 1 may be mounted at a position where the rotation angle at the start of discharge of the first stage compressor 1 is delayed by 60 ° with respect to the rotation angle at which the suction of the second stage compressor 2 is completed.

FIG. 23 shows the discharge volume gradient of the first stage
It is the figure which overlapped and drawn the suction volume gradient of the stage side compressor 2. In FIG. 23, the rotation angle of the male rotor on the horizontal axis is based on the rotation angle of the male rotor 4 of the two-stage compressor 2. In FIG. 23, the rotation angle of the male rotor is-
In the range of 120 ° to -80 ° and -60 ° to 40 °, the intermediate pressure is not stable because the discharge volume gradient of the first stage compressor is larger than the suction volume gradient of the second stage compressor. Become. Conversely, when the rotation angle is in the range of -80 ° to -60 °, the rotation angle is 2
Since the suction volume gradient of the stage-side compressor is large, the intermediate pressure becomes unstable and becomes low. The tendency of the intermediate pressure to fluctuate with respect to the rotation angle of the male rotor is also shown in FIG.

When the number of teeth of the male rotor is three, the cycle of these series of changes is repeated three times per rotation, that is, the rotation angle of the male rotor repeats every 120 °. Thus, with respect to the rotation angle of the male rotor, the stable intermediate pressure when the intermediate pressure is equal to the discharge volume gradient of the first stage compressor and the suction volume gradient of the second stage compressor is:
It will fluctuate on both the positive side and the negative side.

[Problem to be Solved by the Invention] In the above-described conventional technology, the phase difference between the rotation angle of the male rotor of the one-stage compressor and the rotation angle of the male rotor of the two-stage compressor is
If it is not installed in a certain position, the following problems occur.

(1) Since the intermediate pressure fluctuates, the gas torque that the rotor receives due to the fluctuation of the intermediate pressure occurs in both the first and second stage compressors. For this reason, the vibration of the rotor increases, and the noise due to the rattling noise on the tooth surface of the rotor increases.

(2) Since the discharge pressure of the first-stage compressor and the suction pressure of the second-stage compressor fluctuate, the performance deteriorates due to an insufficient discharge air amount, excessive shaft power, etc. of the entire compressor.

(3) The discharge pressure also fluctuates due to fluctuations in the suction pressure of the two-stage compressor. For this reason, it becomes impossible to maintain a stable discharge pressure, which adversely affects equipment at the compressor outlet.

An object of the present invention is to provide a multi-stage screw compressor in which the fluctuation pressure of the intermediate pressure is extremely small and a stable intermediate pressure can be obtained.

[Means for Solving the Problems] A multistage in which a low-pressure side compressor and a high-pressure side compressor are connected, a fluid compressed by the low-pressure side compressor is sucked into the high-pressure side compressor, compressed to a predetermined pressure and discharged. In the screw compressor, the male rotor of the low-pressure side compressor and the male rotor of the high-pressure side compressor are connected so that the discharge volume gradient of the low-pressure side compressor is larger than the suction volume gradient of the high-pressure side compressor. The discharge start and the suction completion timing of the high-pressure side compressor were connected in synchronization.

[Operation] The discharge volume gradient of the low-pressure side compressor is always larger than the suction volume gradient of the high-pressure side compressor for any rotation angle of the male rotor. For this reason, the fluctuation pressure at the intermediate pressure becomes extremely small. Further, a fluctuating pressure is generated only on the plus side with respect to the stable intermediate pressure, and no fluctuating pressure on the minus side is generated.

As a result, the fluctuation of the gas torque of the rotor caused by the fluctuation of the intermediate pressure can be mitigated, so that the vibration of the rotor can be reduced, and the noise due to the rattle on the tooth surface of the rotor can be reduced. In addition, since fluctuations in the discharge pressure of the low-pressure compressor and the suction pressure of the high-pressure compressor can be reduced,
The overall performance of the compressor can be improved. further,
Fluctuations in the suction pressure of the high-pressure compressor can be mitigated, so that fluctuations in the discharge pressure can be reduced, thereby reducing adverse effects on other equipment at the compressor outlet caused by fluctuations in the discharge pressure. Can be eliminated.

Embodiment One embodiment of the present invention will be described below with reference to FIGS. 1 to 12.

FIG. 1 is a longitudinal sectional view of a two-stage screw compressor. As shown in FIG. 1, a first-stage compressor 1 as a low-pressure compressor and a two-stage compressor 2 as a high-pressure compressor are each composed of a pair of a male rotor and a female rotor. Are supported by bearings 6 to 9 provided in the casing and mesh with each other to rotate.

The male rotor 3 of the first-stage compressor 1 and the male rotor 4 of the second-stage compressor 2 are connected at their shaft ends by a gear coupling 11, and the motor 14 drives the first-stage compressor 1. Accordingly, the first-stage and second-stage compressors 1 and 2 respectively rotate.

A delivery casing 10 is attached to the discharge side of the two-stage compressor 2.

The fluid sucked from the suction port 13 of the first-stage compressor 1 is compressed to an intermediate pressure and discharged from the discharge port 12 of the first-stage compressor 1, passes through the intermediate casing 5, and passes through the intermediate casing 5. , Where it is compressed to a predetermined pressure and discharged.

FIG. 2 (A) is an enlarged longitudinal sectional view of a connecting portion of the male rotor 3 of the first-stage compressor 1 and the male rotor 4 of the second-stage compressor 2,
Figure (B) is a II _B -II _B line sectional view of FIG. 2 (A). As shown in FIGS. 1 and 2 (A), the male rotor 3 of the first-stage compressor 1 and the male rotor 4 of the second-stage compressor 2
They are connected by a gear coupling 11. The two male rotors 3 and 4 are connected to each other via a pin 15 as a positioning coupling means so that the timing of the start of discharge of the first-stage compressor 1 and the completion of suction of the second-stage compressor 2 are matched. The phase of the rotation angle of the male rotor 3 of the compressor 1 and the phase of the rotation angle of the male rotor 4 of the two-stage compressor 2 are combined. As a result, during the operation of the two-stage screw compressor, the moment of the opening of the discharge port 12 of the first-stage compressor 1 and the opening of the suction port (see reference numeral 18 in FIG. 4) of the two-stage compressor 2 are always maintained. You can synchronize with the moment.

FIG. 3 is a cross-sectional view (cross-sectional view taken along the line III-III in FIG. 1) of the discharge port portion of the first-stage compressor 1, and FIG. 4 is a cross-sectional view of the second-stage compressor 2 near the suction port. Cross-sectional view (IV-IV in FIG. 1)
FIG. FIG. 5 shows the male rotor 3 in FIG.
Helix outer periphery from the intersection P ₁ of the outer periphery of the female rotor 16 P along the outer circumference of the male rotor 3 'to _1, also from the intersection point P ₁ P along the outer circumference of the male rotor 16' until ₁ developed in the plane of the It is a development view. Several lines obliquely shown in FIG. 5 indicate lines along the torsion direction of the tooth tips of the male rotor 3 and the female rotor 16. This line moves from the suction side to the discharge side in the direction indicated by the arrow in FIG. 5 as the rotor rotates. Third
Endpoint m ₁ and f ₁ of the axial direction of the discharge port 12 in the figure correspond to the point m ₁ and f ₁ in FIG. 5, respectively.

In FIGS. 3 and 5, when the lines at the tooth tips of the male rotor 3 and the female rotor 16 reach the points m ₁ and f ₁ , respectively, it is the moment when the discharge port 12 is opened.
And the fluid compressed between the tooth spaces of the female rotor 16 and the discharge port 12
Is discharged to

FIG. 6 shows the outer periphery of the male rotor 4 and the female rotor in FIG.
P along the outer periphery from the intersection P ₂ of the male rotor 4 and the outer circumference of the 17 '
₂ until, also P along the intersection P ₂ on the outer circumference of the female rotor 17 '
₂ is a development view of a helix in which up to ₂ is developed on a plane. Fifth
Similarly to the drawing, the line at the tip of the rotor moves from the suction side to the discharge side in the direction indicated by the arrow in FIG. 6 with the rotation of the rotor. The points m ₂ and f _{2 in} FIG. 6 indicate the end point of the suction port. When the tip of the rotor reaches the point m ₂ and f ₂ , it is the moment when the suction port closes and the suction ends. .

FIG. 7 is a volume curve showing the degree to which the volume between one tooth groove of meshing between the male rotor and the female rotor changes with respect to the rotation angle of the male rotor on the suction side and the discharge side. As shown in FIG. 7, the volume changes linearly with respect to the rotation angle of the male rotor in the range from −300 ° to −60 ° on the suction side and from 60 ° to 300 ° on the discharge side. . It does not change linearly in other rotation ranges.

FIG. 8 is a diagram showing the change of the suction volume with respect to the rotation angle of the male rotor of the two-stage compressor in the case of the rotor showing the volume curve of FIG. Here, at the suction port, the volume between the tooth spaces of three teeth of the male rotor is synthesized, so that the suction volume V on the vertical axis in FIG. It is a value synthesized by shifting by ゜. It can be seen from FIG. 8 that the suction volume does not have a linear characteristic with respect to the rotation angle θ of the male rotor.

FIG. 9 is a diagram showing how the value of the change amount dv of the suction volume (= dv / dθ) with respect to the change dv / dθ of the rotation angle of the male rotor in FIG. 8 changes depending on the rotation angle of the male rotor. It is. This dv / dθ is called the suction volume gradient.
That is, this is equivalent to the amount of air taken in per unit time because the rotational angular velocity of the male rotor is constant.

From FIG. 9, it can be seen that the rotation angle of the male rotor is from -120 ° to -6.
In the range up to 0 °, the suction volume gradient increases as the rotation proceeds. In other words, the amount of air taken in per unit time increases. Conversely, when the rotation angle of the male rotor is in the range of -60 ° to 0 °, the gradient of the suction volume decreases as the rotation proceeds. In other words,
The amount of air taken in per unit time falls as the rotation progresses. The rotation angle of the male rotor in FIGS. 8 and 9 is 0 ° when the suction is completed.

FIG. 10 is a diagram showing a change in the discharge volume with respect to the rotation angle of the male rotor of the one-stage compressor in the case of the rotor having the characteristic curve of FIG. From FIG. 10, when the rotation angle of the male rotor at the start of discharge is set to 120 °, the discharge volume changes directly when the rotation angle of the male rotor is in the range of 0 to 90 °. It can be seen that there is no linear change in the range.

FIG. 11 is a diagram showing how the value of the change amount dv of the discharge volume (= dv / dθ) with respect to the change amount dθ of the rotation angle of the male rotor in FIG. 10 changes depending on the rotation angle of the male rotor. It is. This dv / dθ is referred to as a discharge volume gradient, which corresponds to the amount of air discharged per unit time. From FIG. 11, assuming that the discharge start time is 120 ° as a reference of the rotation angle of the male rotor, the volume gradient is constant in the rotation angle range of 0 to 90 °, but in the range of 90 to 120 °,
That is, before the discharge port opens and discharge starts, 30
It can be seen that the range is not constant but falls down.

The suction volume gradient of the two-stage compressor shown in FIG.
If the gradient of the discharge volume of the first stage compressor shown in FIG. 11 is always equal, the intermediate pressure is constant, but if it is not equal, the intermediate pressure fluctuates due to the volume difference.

FIG. 12 is a drawing in which the discharge volume gradient of the first-stage compressor 1 and the suction volume gradient of the second-stage compressor 2 are superposed in a state where the phases of the male rotors 3 and 4 are matched. In FIG. 12, the rotation angle of the male rotor on the horizontal axis is based on the rotation angle of the two-stage compressor at the completion of suction. In this case, the discharge volume gradient of the first stage compressor 1 is larger than the suction volume gradient of the second stage compressor 2 regardless of the rotational angle of the male rotor. Therefore, the fluctuation pressure of the intermediate pressure caused by the difference in the volume gradient is generated only on the plus side with respect to the stable intermediate pressure when there is no difference in the volume gradient. FIG. 12 also shows the tendency that the intermediate pressure fluctuates with respect to the rotation angle of the male rotor.

In the case of the prior art, as described above, the discharge volume gradient of the first-stage compressor may be larger or smaller than the suction volume gradient of the second-stage compressor depending on the rotation angle of the male rotor. The fluctuating pressure is Δ on the plus side
It occurs on both sides of the [Delta] P ₂ of P ₁ and the minus side and ΔP ₁ + ΔP ₂ is a fluctuating pressure range.

In contrast, the embodiment of the present invention, variable pressure in order to generate only the [Delta] P ₁ of the positive side pressure fluctuation width is [Delta] P _1. Therefore, it is possible to reduce the fluctuating pressure width by ΔP ₂ as compared with the case of the related art.

The positioning and coupling means is not limited to the pin 15, but may be a clamp or the like. In short, any structure may be used as long as the male rotors can be positioned and coupled.

[Effects of the Invention] As described above, according to the present invention, in the multi-stage screw compressor, the male rotor of the low-pressure side compressor and the male rotor of the high-pressure side compressor have a discharge volume gradient of the low-pressure side compressor. Since the start of the discharge of the low-pressure compressor and the timing of the completion of the suction of the high-pressure compressor are connected in synchronization with each other so as to be larger than the suction volume gradient of the high-pressure compressor, the following effects are obtained.

(1) Both the low-pressure side compressor and the high-pressure side compressor can reduce the gas torque fluctuation of the rotor caused by the fluctuation of the intermediate pressure. For this reason, vibration of the rotor can be reduced, and noise due to rattling noise on the tooth surface of the rotor can be reduced.

(2) Fluctuations in the discharge pressure of the low-pressure side compressor and the suction pressure of the high-pressure side compressor can be reduced, so that the performance of the entire compressor can be improved.

(3) Fluctuations in the suction pressure of the high-pressure side compressor can be reduced, so fluctuations in the discharge pressure can be reduced. For this reason, a more stable discharge pressure can be maintained, and the adverse effect on the equipment at the compressor outlet caused by the discharge pressure fluctuation can be eliminated.

[Brief description of the drawings]

FIG. 1 shows an embodiment of the present invention, and is a longitudinal sectional view of a two-stage screw compressor. FIG. 2 (A) shows a male rotor and a high-pressure side compressor of a one-stage side compressor which is a low-pressure side compressor. 2-stage-side enlarged longitudinal sectional view of the connecting portion of the compressor of the male rotor is a machine, II _B -II _B line sectional view of FIG. 2 (B) a second figure (a), FIG. 3 is a first diagram 3 is a sectional view taken along line III-III, FIG. 4 is a sectional view taken along line IV-IV in FIG. 1, FIG. 5 is a helix development view of the rotor in FIG. 3, and FIG. 6 is a helix development view of the rotor in FIG. , FIG. 7 is a volume curve diagram of the intermeshing one tooth space of the rotor, FIG. 8 is a suction volume curve diagram of the two-stage compressor in the present invention,
FIG. 9 is a suction volume gradient diagram of the compressor of FIG.
10 is a discharge volume curve diagram of the first stage compressor of the present invention, FIG. 11 is a discharge volume gradient diagram of the compressor of FIG. 10, and FIG. 12 is a first stage compressor and a second stage side of the present invention. It is a figure which shows the volume gradient diagram of a compressor, and the fluctuation | variation pressure of an intermediate pressure. 13 is a longitudinal sectional view of a conventional two-stage screw compressor, FIG. 14 is a sectional view taken along line XIV-XIV of FIG. 13, and FIG.
13 is a cross-sectional view taken along the line XV-XV, FIG. 16 is a helix development view of the rotor of FIG. 14, FIG. 17 is a helix development view of the rotor of FIG. 15, and FIG. Volume curve diagram, FIG. 19 is a suction volume curve diagram of the conventional two-stage compressor, FIG. 20 is a suction volume gradient diagram of the compressor of FIG. 19, and FIG. 21 is a discharge of the first-stage compressor. Volume curve diagram, FIG. 22 is a discharge volume gradient diagram of the compressor of FIG. 21, and FIG. 23 is a diagram showing a volume gradient diagram of the first-stage compressor and the second-stage compressor and a fluctuation pressure of the intermediate pressure. It is. 1 ... 1st stage compressor which is a low pressure side compressor, 2 ... 2nd stage compressor which is a high pressure side compressor, 3 ... Male rotor of 1st stage compressor, 4 ... 2nd stage compressor Male rotor, 5 ... Intermediate casing, 6-9 ... Bearing, 11 ... Gear coupling, 12
…… Discharge port of 1st stage compressor, 13… Suction port of 1st stage compressor, 14… Motor, 15… Pin as positioning coupling means, 16 …… Female rotor of 1st stage compressor , 17 …… 2
Female rotor of stage compressor, 18 ... Suction port of stage compressor.

Claims (1)

(57) [Claims] 1. A low pressure side compressor and a high pressure side compressor are connected,
In the multi-stage screw compressor that sucks the fluid compressed by the low-pressure side compressor into the high-pressure side compressor, compresses the fluid to a predetermined pressure, and discharges the male rotor of the low-pressure side compressor and the male rotor of the high-pressure side compressor, The discharge start of the low-pressure compressor and the suction completion timing of the high-pressure compressor are linked so that the discharge volume gradient of the low-pressure compressor is greater than the suction volume gradient of the high-pressure compressor. Multi-stage screw compressor.