JP2005195027A - Screw fluid machine and screw gear - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a screw fluid machine capable of preventing a local abnormal rise of temperature when used as a vacuum pump or a compression pump, and to provide a screw gear suitable for the screw of such the screw fluid machine. <P>SOLUTION: The screw fluid machine is provided with a male rotor 101 and a female rotor 102 meshing each other, a casing for housing both rotors, a fluid operation chamber formed of the male and female rotors and the casing, and an inflow port and an outflow port for the fluid provided in the casing to be communicated with one end and the other end of the operation chamber. The helix angle of the screw gear constituting the male and female rotors is continuously changed in association with the advance of the helix. Furthermore, the rolling circumferential length of a pitch cylinder relative to the advance direction of the helix is substantially expressed by a monotonously increasing function in the development of a tooth trace rolling curve on the diameter of the pitch cylinder of the screw gear. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

The present invention relates to a screw fluid machine, that is, a screw-type vacuum pump, a compression pump, a motor, and the like, and more particularly to a screw vacuum pump suitable for a low / medium vacuum region from atmospheric pressure to 10 ⁻⁴ Torr level. The present invention also relates to a screw gear itself suitable for the screw pump or the like.

Conventionally, various types of vacuum pumps such as an oil rotary pump, a roots pump, and a diffusion pump have been used in low and medium vacuum regions. For example, in the semiconductor manufacturing field, a predetermined process is performed by storing a wafer in a vacuumed container. In this process, a vacuum pump is used while supplying an inert gas such as N ₂ gas into the container. Suction is performed to remove impurities (O ₂ , CO _2, etc.) in the container, and a vacuum state of several Torr to 10 ⁻⁴ Torr level is obtained. As a vacuum pump used in such a semiconductor manufacturing process, an oil rotary pump, a roots type mechanical booster pump, or the like is used.

  However, in the oil rotary pump, the lubricating oil used is in contact with various gases (for example, arsenic, gallium, chlorine, Poly-Si, fluorine, etc.) used in the semiconductor manufacturing process, and the life as a lubricating oil is shortened. In addition, there is a problem that oil molecules are contaminated in the semiconductor manufacturing container and are not preferable in the semiconductor manufacturing process.

In addition, since the above-mentioned pump has a narrow pressure range for normal operation, several types of pumps must be switched before reaching a predetermined pressure. One vacuum pump from atmospheric pressure to 10 ⁻⁴ Torr level. There was a problem that the exhaust could not be exhausted.

In order to solve these problems, an oil-free screw vacuum pump disclosed in JP-A-60-216089 has already been proposed.
The screw vacuum pump disclosed in the above publication is a non-lubricated screw vacuum pump that can cover the pressure range with a single unit.

The outline of this screw type vacuum pump is demonstrated based on FIG. As shown in the figure, the male rotor 10 and the female rotor 11 are rotatably supported by bearings 14, 15, 16, and 17 in the main casing 12 and the suction casing 13.
The male rotor 10 and the female rotor 11 are formed by screw gears, and the gears have a constant helical twist angle, and the rotation axis direction pitch and the axis perpendicular plane pitch are also constant. It does not change with the change of the rotation angle.

Further, since the suction side 10a of the rotor has a low pressure of 10 ⁻⁴ Torr level and the discharge side 10b becomes atmospheric pressure, the radial load acting on the rotor is much smaller on the suction side. Therefore, the bearings 14 and 15 on the suction side support the radial load and the thrust load using a deep groove bearing, and the bearings 16 and 17 on the discharge side support only the radial load using a cylindrical roller bearing.
Timing gears 18 and 19 are attached to the shaft ends of the rotor, and the gap between the rotors is adjusted so that the male and female rotors 10 and 11 do not contact each other.
The bearings 14 and 15 are lubricated by droplet lubrication, and the lubricating oil 21 accumulated in the suction cover 20 is splashed by the timing gear.

On the other hand, a disk 22 is attached to the male rotor shaft for lubricating the bearings 16 and 17, and the lubricating oil 24 in the discharge cover 23 is splashed by the disk 22.
The shaft seals 25, 26, 27, and 28 prevent the lubricating oil from the bearings and timing gear from entering the working chamber.

Since the pressure in the discharge side working chamber 10b and the discharge cover 23 of the rotor is almost atmospheric pressure, the differential pressure acting on the shaft seals 27 and 28 on the discharge side is relatively small, but the suction side working chamber 10a has a level of 10 ⁻⁴ Torr. If the inside of the suction cover 20 is opened to the atmosphere because of the pressure, the differential pressure acting on the suction-side shaft seals 25 and 26 increases, making sealing difficult.
Therefore, the inside of the suction cover 20 is communicated with the low-pressure working chamber 10c by the exhaust pressure pipes 29 and 30, and the pressure inside the suction cover 20 is lowered to reduce the differential pressure acting on the shaft seals 25 and 26, thereby enhancing the sealing effect. Yes.

Since the inside of the suction cover 20 is filled with oil droplets, the suction cover is provided with a droplet separation chamber 31 in order to prevent entry into the working chamber through the oil pressure relief pipes 29, 30. An oil trap 32 is attached to the pressure tube.
Even if oil enters the working chamber through the exhaust pressure pipe, the exhaust pressure port 34 of the main casing 12 is connected to the suction chamber 33 so that the oil does not flow back to the suction port 33 side. 33 is opened to a position after being completely closed.

The working chamber 10c of the male rotor 10 has the female rotor 11 and two meshing portions 36 and 37 until the working chamber passes through the suction port 33 and communicates with the discharge port 35. Similarly, the working chamber 10c of the female rotor The working chamber 11c has a male rotor and two meshing portions 38, 37.
Along with the rotation of the rotor, the gas is sucked into the working chamber formed by the rotor tooth groove and the casing from the suction port 33 and discharged from the discharge port 35.
The working chambers 10c and 11c transfer the gas with a constant volume as the rotor rotates, but the working chambers 39 and 40 at the position where the rotor further rotates reduce the volume and compress the gas as the rotor rotates.

Furthermore, the state in which the male rotor 10 and the female rotor 11 are engaged with each other will be described with reference to FIG. 3, which is a schematic view developed in the circumferential direction of the rotor.
As shown in the figure, the casing 12 that covers the rotor has one end in the axial direction that is largely open as a gas suction port 33, and a discharge port 35 is provided on the opposite side.
Except for these two ports, the casing 12 covers the rotors 10 and 11 with a minute gap, and a V-shaped working chamber is formed by the rotor and the casing.

When the rotor rotates, the meshing portion of both rotors moves from the suction port 33 toward the discharge port 35. At this time, the working chamber 41 reduces its volume and compresses the gas in the working chamber.
On the other hand, since the volume of the working chamber 42 is constant, there is no gas compression action and a transfer action.

That is, the male rotor 10 and the female rotor 11 are formed by a rotor and a casing because the helical twisting angle is always a constant angle, and the rotation axis direction pitch and the axis perpendicular plane pitch are also constant. The volume of the V-shaped working chamber 42 is constant.
However, when the rotor rotates and the meshing portion of both rotors moves from the suction port 33 toward the discharge port 35, the volume of the working chamber 41 is reduced by the end plate 12a of the casing 12.
Therefore, the working chamber 41 acts to reduce the volume and compress and transfer the gas in the working chamber, while the working chamber 42 has a constant volume, so there is no gas compressing action and it simply acts to transfer.

In the drawing, the working chamber 43 is discharging gas through the discharge port 35, and each working chamber communicating with the suction port 33 increases its volume with the rotation of the rotor to perform the gas suction action. Yes. Moreover, the screw fluid mechanism as described above is also used as a compression pump, and can also be used as a motor.
JP-A-60-216089

  As described above, the conventional screw fluid machine used as a vacuum pump reduces the volume of the working chamber and compresses the fluid, and makes the volume of the working chamber constant and exerts a compressive action on the fluid. And a working chamber that simply performs a transfer action.

  Therefore, in the conventional screw vacuum pump, the temperature of a part of the rotor and the casing of the vacuum pump abnormally rises due to a local (high compression) pressure increase. That is, as shown by the dotted line in FIG. 8, there is a tendency that the temperature on the discharge side where the working chamber for compressing the gas is reduced and the volume of the working chamber is decreased becomes abnormally high. As a result, due to the local temperature rise, the thermal expansion of the members constituting the screw vacuum pump becomes non-uniform, and the dimensional accuracy between the engagement of the casing and the rotor, the male rotor, and the female rotor cannot be improved. There were technical issues such as.

  The first object of the present invention is to solve such problems, and it is intended to provide a screw fluid machine that does not cause a local abnormal temperature rise when used as a vacuum pump or a compression pump. Is. The present invention also provides a screw gear suitable for the screw of such a screw fluid machine.

The characteristics of the exhaust speed of the conventional screw vacuum pump described above are shown by the dotted line in FIG.
As is apparent from this figure, this conventional screw vacuum pump can achieve an ultimate pressure of 10 ⁻⁴ Torr, but the exhaust speed decreases from 10 ⁻² Torr on the high vacuum side.
Therefore, in order to obtain the ultimate pressure of 10 ⁻⁴ Torr using this conventional screw vacuum pump, there is a technical problem that it takes a considerable time, and it is necessary to shorten this time.

The second object of the present invention is to solve the above-mentioned problems. For example, when used as a vacuum pump, a stable exhaust speed is obtained in an operating range from atmospheric pressure (760 Torr) to 10 ⁻⁴ Torr. It is intended to provide a screw fluid machine that can.

Further, in a conventional screw fluid machine used as a vacuum pump, a male rotor is first rotationally driven by a single motor, and a female rotor is rotated via a timing gear.
Therefore, a load for rotating the female rotor is applied to the timing gear. As a result, when the rotor is rotated at a high speed, there is a technical problem that noise due to the engagement of the timing gear is generated and the working environment is deteriorated.

  The third object of the present invention is to solve such a problem, and is intended to provide a screw fluid machine with low noise even when rotated at high speed.

In another conventional screw vacuum pump, when operating in a state where the suction pressure is close to atmospheric pressure, an excessive increase in the pressure in the working chamber is prevented to prevent an abnormal increase in the temperature of the vacuum pump. As shown in FIG. 4, a pressure adjusting device 27 is provided on the lower surface of the casing 12 and in the axial direction of the rotor.
As shown in FIG. 5, the pressure adjusting device 27 includes a discharge port 52 formed in the lower portion of the casing 12, a valve rod 53 that opens and closes the discharge port 52, a spring 54 that supports the weight of the valve rod 53, The valve box 55 accommodates the valve rod 53 and the spring 54, and the atmosphere opening 56 that discharges the gas discharged from the discharge port 52 formed in the valve box 55 to the outside. An O-ring 57 is attached to the valve stem 53. However, in the pressure adjusting device 50 provided at such a position, the working chamber 51a and the adjacent working chamber 51b may communicate with each other via the discharge port 52 as shown in FIG. The arrows in the figure indicate the gas flow at this time.
That is, since the tooth tip 58 of the rotor does not have a sufficient width, the discharge port 52 extends over the working chamber 51a and the adjacent working chamber 51b.
As a result, there is a problem that gas is leaked from the high-pressure working chamber 51a to the low-pressure working chamber 51b, and extra time is required until the suction side reaches a predetermined degree of vacuum.

  The fourth object of the present invention is to solve such a technical problem, and prevents an increase in shaft torque due to excessive compression, prevents an abnormal temperature rise, A screw vacuum pump in which the pressure on the side is set to a predetermined degree of vacuum in a short time is provided.

  In order to achieve the first object, a screw fluid machine according to the present invention includes a male rotor and a female rotor that mesh with each other, a casing that houses both rotors, and a fluid working chamber formed by the male and female rotors and the casing. And a screw fluid machine comprising a fluid inlet and outlet provided in the casing so as to communicate with one end and the other end of the working chamber. However, the basic configuration is that it is configured to continuously change as the twisting proceeds.

When the fluid machine configured as described above is used as a vacuum pump or a compression pump, the tooth helix angle of the male rotor and the female rotor changes as the torsion progresses. The volume of the V-shaped working chamber is continuously reduced as it proceeds from the inlet to the outlet.
Therefore, the working chamber formed by the male and female rotors and the casing has suction, continuous compression transfer, and discharge actions, and does not have a simple transfer action with a constant volume of the working chamber. Abnormal temperature rise due to the rise can be prevented.
Further, as the screw gear of such a screw fluid machine, the rolling circumference of the pitch cylinder with respect to the torsional traveling direction is a substantially monotonically increasing function in the development diagram of the tooth rolling curve on the pitch cylinder diameter of the screw gear. The screw gear according to the present invention can be used, and as a result, the sealing property at right angles to the rotation axis is improved, and the air tightness of the fluid working chamber can be improved.
In addition, since the screw gear according to the present invention can be used as a normal transmission gear, the torsion angle changes with time as it rotates, so that a load with temporal fluctuation in the axial direction can be effectively applied. Can be processed.

In order to achieve the second object, a screw fluid machine according to the present invention includes a male rotor and a female rotor that mesh with each other, a casing that houses both rotors, and a fluid formed by the male and female rotors and the casing. A screw fluid machine comprising a working chamber and a fluid inlet and outlet provided in the casing so as to communicate with one end and the other end of the working chamber. A root portion is formed at one end of at least one of the screw gear portions.
When the screw fluid machine configured as described above is used as a vacuum pump, for example, in addition to the pumping action by the screw part due to the rotation of the male and female rotors, the screw fluid machine is also pumped by the root part provided on at least one end side of the screw part. Since the operation is performed, a stable exhaust speed can be obtained in the operating range from atmospheric pressure (760 Torr) to 10 ⁻⁴ Torr without reducing the exhaust speed in the range of 10 ⁻² Torr to 10 ⁻⁴ Torr. . Further, when used as a compressor, a high discharge pressure can be obtained.

  In order to achieve the third object, a screw fluid machine according to the present invention includes a male screw rotor and a female screw rotor that mesh with each other, a casing that houses both rotors, and a fluid formed by the male and female rotors and the casing. In a screw fluid machine comprising a working chamber and a fluid inlet and outlet provided in the casing so as to communicate with one end and the other end of the working chamber, for driving the male and female rotors Each of the rotors, an inverter for sending a driving AC signal or a driving pulse signal to each of the motors, and a controller for sending a control signal for controlling the frequency of the inverter. And the number of rotations of the female rotor is controlled.

In the screw fluid machine configured as described above, when a control signal corresponding to a predetermined number of revolutions, that is, a control signal for controlling the frequency of the inverter is sent from the controller to the inverter, the inverter receives the control signal in response to the control signal. A driving AC signal or a driving pulse signal having a predetermined frequency (reference frequency) is transmitted, and the motors M ₁ and M ₂ are driven at a predetermined number of revolutions according to the driving AC signal or the driving pulse signal. .
Therefore, since the male and female rotors are driven to rotate synchronously by the motors M ₁ and M ₂ , even if the male and female rotors are rotated at a high speed, fluctuations in the rotational speed of the motor are small and the load applied to the timing gear is small. Noise due to gear meshing can be suppressed.

  In order to achieve the fourth object, a screw fluid machine according to the present invention forms a working chamber by a male screw rotor and a female screw rotor that mesh with each other, and a casing that houses both rotors, and is confined in the working chamber. In the screw fluid machine provided with a pressure adjusting device for discharging the generated suction gas from the discharge port with rotation of the rotor and limiting the pressure so that the pressure in the working chamber does not rise above atmospheric pressure, The adjusting device includes: a discharge port formed in a screw end face plate that constitutes a part of the casing; and a discharge valve that is provided outside the discharge port and is opened when the working chamber pressure exceeds the vicinity of atmospheric pressure. A rotor tooth end surface that opens and closes the discharge port, and the rotor tooth end surface is positioned at the discharge port in accordance with the rotation of the rotor. Tooth end face is configured so as to close the inside of the discharge port.

In the screw fluid machine configured as described above, when the pressure of the suction gas is low and the pressure of the working chamber is lower than the vicinity of the atmospheric pressure, the discharge valve of the pressure adjusting device closes the outside of the discharge port.
At this time, since the inside of the discharge port is configured to be closed by the tooth end surface of the screw gear constituting the rotor, even if the rotor rotates, the working chamber does not communicate with the adjacent working chamber, There is no need for extra time until gas leaks from the high working chamber to the low working chamber and the suction side reaches a predetermined degree of vacuum.
When the pressure of the suction gas is high and the pressure in the working chamber is higher than the vicinity of the atmospheric pressure, the discharge valve of the pressure regulator is released, and the gas in the working chamber is released to the outside through the discharge port.
When the suction pressure decreases and the working chamber pressure does not reach atmospheric pressure immediately before the working chamber communicates with the discharge port, all the discharge ports of the pressure regulator are closed, and the gas in the working chamber is Without being discharged from the pressure adjusting device to the outside, it is discharged from the discharge port.

As is clear from the above description, according to the screw fluid machine according to the present invention, the tooth helix angle of the male rotor and the female rotor is changed according to the progress of the torsion, and therefore the V formed by the rotor and the casing. The volume of each fluid working chamber in the shape of a letter can be continuously increased or decreased as the rotor rotates.
As a result, it is possible to suppress an abnormal rise in temperature locally, and there is an effect that the dimensional accuracy between the engagement of the casing and the rotor, the male row, and the female rotor can be improved. Is.
In addition, as the torsional gear of such a screw fluid machine, the rolling circumference of the pitch cylinder with respect to the direction of twisting in the development of the tooth rolling curve on the pitch cylinder diameter of the screw gear is substantially a monotonically increasing function. The screw gear according to the present invention can be used, and as a result, the sealing property at right angles to the rotation axis is improved, and the air tightness of the fluid working chamber can be improved.
In addition, since the screw gear according to the present invention can be used as a normal transmission gear, the torsion angle changes with time as it rotates, so that a load with temporal fluctuation in the axial direction can be effectively applied. Can be processed.

Further, according to the fluid machine according to the present invention, since the root portion is provided on at least one end side of the screw portion of the male and female rotors, for example, when used as a vacuum pump, the exhaust speed is greatly improved, and A vacuum pump can efficiently obtain a stable exhaust speed from atmospheric pressure to a medium vacuum region of 10 ⁻⁴ Torr, and can provide a high discharge pressure when used as a compression pump. .

  Further, according to the fluid machine of the present invention, since the male and female rotors are driven synchronously, there is an effect that noise due to the engagement of the timing gear can be suppressed even when the rotor is rotated at a high speed.

Furthermore, the fluid machine according to the present invention is configured such that the inside of the discharge port is closed by the tooth end surface of the rotor while the tooth end surface of the rotor is positioned at the discharge port according to the rotation of the rotor. There is no communication between one working chamber and another adjacent working chamber by the outlet.
As a result, there is an effect that gas does not leak from one working chamber having a high predetermined pressure to another working chamber having a low pressure, and no extra time is required until the suction side reaches a predetermined degree of vacuum.

  In addition, since the pressure in the working chamber can be suppressed to almost the atmospheric pressure or less, even when the operation is performed in a state where the suction pressure is close to the atmospheric pressure, an increase in shaft torque due to excessive compression can be prevented, and power consumption can be suppressed. In addition, since there is no compression more than necessary, the temperature of the screw vacuum pump does not rise abnormally, and the dimensional accuracy between the engagement of the casing and the rotor, the male rotor and the female rotor, etc. can be improved. This is an effect.

First, in the screw fluid machine according to the present invention, one example of a vacuum pump in the case where the screw has a continuously changing helix angle is also illustrated in conjunction with the description of the embodiment of the screw gear (screw) according to the present invention. This will be described with reference to FIGS.
The present inventors have eliminated working chambers with a constant volume that merely perform a transfer action without exerting a compressing action on the sucked gas, and continuously reducing the volumes of all working chambers, Focused on doing.
Then, in order to continuously reduce the volume of the working chamber, the tooth helix angle of the gear constituting the male rotor and female rotor of the screw vacuum pump is changed according to the rotation angle of the rotor, and is formed by the rotor and the casing. The V-shaped working chamber volume is to be changed.
Therefore, since the important point here is the shape of the screw gears constituting the male rotor and the female rotor, in one embodiment described below, the shape of the screw gears of the screw vacuum pump will be described. Other structures are the same as those in the prior art, and a description thereof is omitted.

A screw gear used in the screw vacuum pump according to the present embodiment will be described with reference to FIGS.
Here, FIG. 6 is a plan view of the screw gear, and FIG. 7 shows the rolling circumferences x _M and x _F of the male and female gears on the pitch cylinder on the horizontal axis and the travel amount y in the rotational axis on the vertical axis. shows the development of the tooth-trace rolling curve of the female gear male gear, x _F -y plane in the x _M -y plane.
In this x in those for the left half of the female gear figure represent x _F, x in that for the right half of the male gear represents x _M. The sign of x is assumed to move from the suction side to the discharge side when the tooth trace of the gear is traced in the positive direction. That is, the male gear has a positive direction in the right direction, and the female gear has a positive direction in the left direction. The male gear is a gear used for the male rotor, and the female gear is a gear used for the female rotor.
In FIG. 7, the position y of the suction port of the rotor is set to 0, and the position of the discharge port is set to y = L. In the male and female rotors, the male and female teeth on the pitch cylinder coincide with each other at the suction port (y = 0), and the point is x _M = x _F = 0. Here, the tooth rolling curve is generally also called a helical winding.

Moreover, there is no restriction | limiting about the effective range of x of FIG. That is, for x,
x _M ≧ 0, x _F ≧ 0
Is an effective range, and y is determined by the length L of the rotor, and the range of y is 0 ≦ y ≦ L
Given in.

Development view of the tooth rolling curve of the male and female rotors extending from the point where the male and female rotors are in contact with each other on the pitch cylinder (ie, x _M = 0 and x _F = 0) at the suction port (y = 0). 7 originates from the origin in FIG. 7, and both y increase as x increases. That is, for the male rotor, y is a monotonically increasing function of x _M, the female rotor, y is a monotonically increasing function of x _F.
This does not change even if x and y are exchanged and y is regarded as an independent variable and x is a function of y. That is,
For male rotor: x _M = F _M (y) (1)
Is a monotonically increasing function of y,
For female rotors x _F = F _F (y) (2)
Is a monotonically increasing function of y.
Also, since both pass through the origin,
F _M (0) = F _F (0) = 0 (3)
It is.

Here, β _Mg , β _Fg , θ _M, θ _F by β _Mg ; Torsion angle of male rotor on pitch cylinder β _Fg ; Torsion angle of female rotor on pitch cylinder θ _M ; Male rotor rotation angle θ _F Representing the rotation angle of the female rotor. The twist angles β _Mg and β _Fg are angles shown in FIG. Also, if the rotation angles θ _M and θ _F are the radius of the male and female pitch cylinders R _M and R _F ,
θ _M = x _M / R _M (4)
θ _F = x _F / R _F (5)
It is represented by

Using the equations (1) and (2), the torsion angles β _Mg and β _Fg of the male and female rotors are tan β _Mg = dx _M / dy = dF _M / dy (6)
tan β _Fg = dx _F / dy = dF _F / dy (7)
Given in.

The rotor helix angle is continuously increased so that each fluid working chamber defined by the engagement of the male and female screw rotors moves in the vacuum pump discharge direction while continuously reducing the volume as the rotor rotates. Let This is equivalent to continuously increasing dF _M / dy and dF _F / dy from the equations (6) and (7). That is, F _M (y) and F _F (y) given by equations (1) and (2) both pass through the origin, are monotonically increasing functions with respect to y, and their differential coefficients are also monotonically increasing functions. . That is, the functions F _M (y) and F _F (y) are in the domain 0 ≦ y ≦ L of y,
F _M (0) = 0, F _F (0) = 0 (8)
dF _M (y) / dy> 0, dF _F (y) / dy> 0 (9)
d ² F _M (y) / dy ² > 0, d ² F _F (y) / dy ² > 0 (10)
Must be satisfied. That is, an arbitrary function satisfying the expressions (8), (9), and (10): x _M = F _M (y), x _F = F _F (y)
Can be adopted as a developed view of the male and female rotor tooth rolling curve.

As a condition for the engagement of the male and female rotors, the male tooth twist angle and the female tooth twist angle on the pitch cylinder must be equal in size and in opposite directions. However, in the analysis so far, since the positive directions of the rolling circumferences x _M and x _F of the male and female rotors on the pitch cylinder are opposite to each other, the meshing conditions of the male and female rotors are all y values. so,
β _Mg = β _Fg (11)
Must be met. From this condition,
tan β _Mg = tan β _Fg (12)
That is, from the expressions (6) and (7), dx _M / dy = dx _F / dy (13) for all y values in the domain.
The following conditions are obtained.

From the expressions (12) and (13), it is concluded that the functions of x _M = F _M (y) and x _F = F _F (y) are exactly the same. That is, it is concluded that the curve shown in FIG. 7 is symmetrical with respect to the y-axis. In other words, in designing the torsional angle change type rotor,
F (0) = 0, dF / dy> 0, d ² F / dy ² > 0 (14)
An arbitrary function F (y) that satisfies the following is selected, so that x _M = F _M (y), x _F = F _F (y) (15)
And

Equal axial plane perpendicular pitch T on the pitch cylinder, if the number of teeth of the male gear N _M, the number of teeth of the female gear and N _F,
T = 2πR _M / N _M = 2πR _F / N _F (16)
The development diagram of the tooth rolling curve of the rotor corresponding to another tooth type is obtained by translating x = F (y) by mT in the x-axis direction. However, m is a positive or negative integer. In FIG. 7, these curves are indicated by dotted lines.

As the simplest example, the following quadratic function as F (y):
F (y) = Ay ² + By (A> 0, B> 0) (17)
Can be selected. The curve shown in FIG. 7 is an example of such a quadratic curve.

In the torsional angle change type gear specified as described above, it is assumed that the developed view of the tooth rolling curve on the pitch cylinder is given by an arbitrary function satisfying the equation (14). Based on the gradient change, the tooth helix torsion angle on the pitch cylinder is changed corresponding to the rotation of the gear, and on the basis of this, the tooth profile is changed to the helical helix angle of a known helical gear or screw gear. Based on the basic idea, meshing is performed by matching the axis perpendicular pitch T on the plane perpendicular to the rotation axis on the pitch cylinder, and the rotation axis direction pitch t _s changes every moment as the rotation angle changes. The twist advances in the direction of the rotation axis (y direction) while maintaining the meshing state and the tooth profile on the plane perpendicular to the rotation axis.

That is, since the rolling circumference on the pitch cylinder and the amount of twisting direction are the same for the male and female rotors, the length of the winding on the pitch cylinder of the male and female rotors is also equal. That is, in any domain of y [y _i , y _j ],

Thus, corresponding to the meshing of both gears, the lengths of the windings on the pitch cylinder in the domain [y _i , y _j ] are equal.
The tooth rolling curve is also expressed as a function of the rotation angle, and the rotation angle and the amount of tooth rolling are in a proportional relationship. The diameters of the male and female tooth-shaped portions other than the pitch circle diameter, the lengths of the helical windings at R _M ′ and R _F ′ are calculated using the equations (4) and (5), and x _M and x _F in the equation (A). X ' _M = x _M R _M ' / R _M x ' _F = x _F R _F ' / R _F
Is obtained by replacing with Therefore, the equation corresponding to (A) does not hold for contact portions having a diameter other than the pitch cylinder, and is adjusted by slipping. That is,

となる。

It becomes.

In order to engage the male and female rotors, θ _M N _F = θ _F N _M (18) between the rotation angles θ _M and θ _F.
Is necessary. Here, N _M and N _F are the numbers of teeth of the male and female rotors, respectively. Further, the radii R _M and R _F of the pitch cylinder are R _M N _F = R _F N _M (19) due to the nature of the pitch cylinder.
There is a relationship. If θ _M and θ _F change while maintaining the equation (18), y _M (θ _M ) = y _F (θ _F ) (20)
Holds.

Further, the axial progression of Omesuha streaks y _M (theta _M), may from y _F (theta _F), is the rotation axis direction pitch t _s theta (theta from (20), even theta _M even theta _F Not as a function). Although t _s changes as θ increases, pitches t _V− and t _{V +} around the position of y (θ) are
t _V− = y _M (θ _M ) −y _M (θ _M −2π / N _M )
= Y _F (θ _F ) −y _F (θ _F −2π / N _F )
t _{V +} = y _M (θ _M + 2π / N _M ) −y _M (θ _M )
= Y _F (θ _F + 2π / N _F ) −y _F (θ _F ) (21)
Given in.

Therefore, the pitches t _sg and t _s (= t _sg ) in FIG. 7 indicate the pitches at the meshing portions of both rotors, and t _sg (n, n + 1) and t _s (n, n + 1) are t _sg (n, n + 1). ) = Y _M {2π (n + 1) / N _M } −y _M (2πn / N _M )
t _s (n, n + 1) = y _F {2π (n + 1) / N _F } −y _F (2πn / N _F )
... (22)
It is. The increase rate dy / dθ of y (θ) is dy / dθ = Rdy / dx = R / (dx / dy) = R / (dF / dy)
Therefore, the increase rate of y (θ) is inversely proportional to dF / dy, that is, the increase rate gradually decreases as y increases. This means that the pitch in the rotation axis direction gradually decreases as y increases, so that t _s (n−1, n)> t _s (n, n + 1), t _sg (n−1, n)> t _sg (n , N + 1). On the other hand, since the axis perpendicular plane pitch does not change, the same tooth profile always appears with rotation.
That is, the volume in a sealed state between the tooth profile of the male gear and the tooth profile of the female gear can be reduced in time with the movement accompanying the rotation.

In the variable torsion angle gear specified above, the tooth rolling curve on the meshing pitch cylinder monotonously changes its slope and changes monotonically as an increasing function. Based on the gradient change, the variable tooth helix torsion angle on the pitch cylinder is defined, and based on this, the tooth profile is determined based on the basic concept of the helical gear and the tooth helix helix angle of the screw gear. Engagement is carried out by matching the axis perpendicular plane pitch T on the plane perpendicular to the rotational axis in the direction of rotation and twisting, and the rotational axis perpendicular pitch t _sg changes with the change in rotational angle. Since the torsion proceeds in the rotation axis direction Y (閘) while maintaining the meshing state on the plane and the tooth shape, the rotation angle and the amount of rolling of the tooth trace have a fixed relationship, and the tooth shape of a pair of male and female gears Match the shape on the plane perpendicular to the axis of rotation So it can be, the rotation axis perpendicular spatial plane n (n _M or n _F) of sequentially appearing with rotation of the rotary axis-th rotation initially identical teeth appear in.
That is, according to this gear, not only has a characteristic as a normal gear, but also has a characteristic as a screw having high sealing performance on a plane perpendicular to the rotation axis.

It is also possible to change the rotation axis direction pitch periodically and continuously.
Therefore, if a male-female rotor is constructed using this gear, the helical twisting angle of the male and female rotors changes according to the rotation angle of the rotor, and as a result, a V-shaped working chamber formed by the rotor and the casing. The volume can be changed continuously.
That is, all the working chambers can be working chambers whose volumes are reduced.

As described above, if the screw vacuum pump or the compression pump is configured using the gear, the volume of the working chamber is continuously changed and continuously compressed and transferred as shown by the solid line in FIG. The temperature of these pumps gradually increases from the suction side toward the discharge side, and does not increase locally.
In addition, the working chamber performs a suction action of sucking gas in a state where it is in communication with the suction port, a continuous compression transfer action of continuously compressing and transferring the gas in the working chamber, and a discharge action of discharging gas in a state of being connected to the discharge port. Since there is no transfer action, it can be driven efficiently.

  Furthermore, since the pitch in the direction of the rotation axis changes, the overall length of the rotor can be shortened compared with a conventional screw fluid machine having an equal pitch, and the screw fluid machine can be reduced in size.

Next, in the fluid machine according to the present invention, an embodiment as a vacuum pump in the case where a root portion is provided on at least one end side of each screw portion of the male and female rotors will be described with reference to FIGS.
FIG. 9 is a perspective view of the male and female rotors used in this embodiment, and FIG. 10 is a plan view of the male and female rotors. 11 is a cross-sectional view of a slew vacuum pump using the male and female rotors shown in FIGS. 9 and 10, and FIG.

First, the characteristics of the present embodiment will be described. The conventional male and female rotors have been formed with a so-called single screw gear, whereas the present embodiment has the male and female rotors formed with the screw gear and roots. There is a feature.
That is, as shown in FIGS. 9 and 10, the male and female rotors 101 and 102 are constituted by screw gear portions 101a and 102a, male root portions 103 and 105, female root portions 104 and 106, and the male roots. The parts 103 and 105 and the female root parts 104 and 106 are formed at both ends of the screw gear parts 101a and 102a.

Also, working chambers 101b and 102b formed by the screw gear portions 101a and 102a of the male and female rotors 101 and 102 and the casing, a male roots portion 103, a working chamber 103a formed by the female roots portion 104 and the casing, Similarly, the operation chambers 105a and 106a formed by the working chambers 101b and 102b, the male root portion 105, the female root portion 106, and the casing are communicated with each other.
Note that rotary shafts 107 and 108 are formed at one end of the male and female rotors 101 and 102, respectively.

Next, a state in which the male and female rotors 101 and 102 are arranged in the casing will be described with reference to FIGS.
As shown in the figure, the male rotor 101 and the female rotor 102 are housed in a main casing 109 and are attached to bearings 111 and 112 and a sub casing 117 attached to an end plate 110 that seals one end surface of the main casing 109. The bearings 118 and 119 are rotatably supported.
On the end plate 110 side of the main casing 109, there is provided a discharge port 109b for discharging the gas compressed by the male and female rotors 101 and 102 to the outside. Further, seal materials 113 and 114 are attached to the bearings 111 and 112, respectively, and the seal materials 113 and 114 prevent lubricating oil from timing gears 115 and 116 described later from entering the working chamber.

Timing gears 115 and 116 housed in the sub casing 117 are attached to the rotating shafts 107 and 108 of the male and female rotors 101 and 102, and the distance between the two rotors is adjusted so that the male and female rotors do not contact each other. Yes.
The bearings 111 and 112 are lubricated by refueling, and the lubricating oil (not shown) accumulated in the auxiliary casing 117 is splashed by the timing gears 115 and 116.
A sub casing 120 is attached to the other end side of the main casing 109. A suction port 109 a is provided on the other end side of the main casing 109.

The screw vacuum pump configured as described above has a working chamber 103a in which gas is formed by the male root part 105, the female root part 106 and the casing from the suction port 109a as the male and female rotors 101 and 102 rotate. It is sucked into 104a. During the suction, the sucked gas is compressed by the working chambers 103a and 104a of the root parts 103 and 104.
Then, it is transferred to working chambers 101b and 102b formed by the screw gear portions 101a and 102a communicating with the working chambers 103a and 104a and the casing. The working chambers 101b and 102b transfer gas while the initial volume is constant as the rotors 101 and 102 rotate, but when the rotor further rotates, the volume is reduced and the gas is compressed.
Further, the compressed gas is transferred to the working chambers 105a and 106a of the male root portion 105 and the female root portion 106 communicating with the working chambers 101b and 102b, and discharged from the discharge port 109b while being compressed. Is done.

  Since the temperature rises due to gas compression outside the main casing 109, a cooling jacket 121 is provided, and cooling water is passed through the jacket to cool the casing 109 and the compressed gas.

As described above, according to the present embodiment, the functions of the screw pump and the roots pump are combined, so that the exhaust speed of the screw vacuum air pump is greatly improved as shown by the solid line in FIG. It is possible to obtain a substantially stable exhaust speed from the atmospheric pressure (760 Torr) to the middle vacuum region of 10 ⁻⁴ Torr and cover a wide operating range. Moreover, when the pump of the said Example is used as a compressor, a high discharge pressure can be obtained.
In the above embodiment, the roots are formed on both ends of the screw gear, that is, on both the suction port side and the discharge port side, but may be formed on only one of them if necessary.
In the above embodiment, the torsion angle of the screw gear may be continuously changed as described in FIGS. 6 and 7, or may be constant as in FIGS. Good.

Next, an embodiment as a vacuum pump when performing synchronous rotation control of the male and female rotors in the fluid machine according to the present invention will be described with reference to FIGS.
In the screw vacuum pump according to this embodiment, the roots are not provided in the male and female rotors 201 and 202 and the motors M ₁ and M ₂ are attached to the rotating shafts 207 and 208 of the male and female rotors 201 and 202. Except for this point, it basically has the same structure as the vacuum pump shown in FIGS.
As shown in FIG. 16, the control circuits for the motors M ₁ and M ₂ are connected to inverters 202 and 203 for sending drive AC signals or drive pulse signals, and the inverters 202 and 203 are control signals for frequency control. Is connected to the controller 204 for sending the message.
That is, when a control signal corresponding to a predetermined number of revolutions is sent from the controller 204 to the inverters 202 and 203, a driving AC signal or a driving pulse signal having a reference frequency corresponding to the control signal is sent to the inverters 202 and 203. The motors M ₁ and M ₂ are driven at a predetermined rotational speed.

The operation of the screw vacuum pump configured as described above will be described. As described above, a control signal corresponding to a predetermined number of revolutions, that is, a control signal for controlling the frequency of the inverters 202 and 203 is sent from the controller 204 to the inverters 202 and 203. Is done.
In response to this control signal, the inverters 202 and 203 send driving AC signals or driving pulse signals having a predetermined frequency (reference frequency) corresponding to the control signals to the motors M ₁ and M ₂ .
In response to the driving AC signal or driving pulse signal, the motors M ₁ and M ₂ are driven at a predetermined rotational speed.

Here, there is no error in the driving AC signal or driving pulse signal sent from the inverters 202 and 203 of the male and female rotors 101 and 102, and the male and female rotors 101 and 102 are sent at a predetermined frequency (reference frequency). The rotations of the female rotors 101 and 102 are synchronized, the male and female rotors 101 and 102 are driven at the same rotational speed, and no load is applied to the timing gears 115 and 116.
As a result, even if the male and female rotors 101 and 102 are rotationally driven at high speed, no load is applied to the timing gears 115 and 116, so that noise due to the engagement of the timing gears can be suppressed.

  In general, the frequency error of the inverter is 0.2 to 0.3%. The frequency error of the inverter makes it impossible to synchronize the rotation of the male and female rotors 101 and 102, and a load for rotating the male and female rotors 101 and 102 is applied to the timing gears 115 and 116. Since the load is small as compared with the above, the noise generated from the timing gears 115 and 116 can be suppressed as compared with the prior art. In addition, since the tooth surface pressure of the timing gear is smaller than that in the prior art, the pump can be operated at high speed, and the pump exhaust speed can be improved or the pump can be downsized.

Next, another example of the motor control system is shown in FIG.
The same members as those shown in FIG.
Also in the case shown in FIG. 15, the motors M ₁ and M ₂ are connected to inverters 202 and 203 for sending driving AC signals or driving pulse signals as in the case shown in FIG. Is connected to a controller 204 that sends out a control signal for controlling the frequency of the inverters 202 and 203. In the motor control system, feedback 205 and 206 circuits for taking in the drive AC signals or drive pulse signals sent from the inverters 202 and 203 are provided. The feedback circuits 205 and 206 are connected to the inverters 202 and 203, respectively. A control signal is sent to

That is, when a control signal corresponding to a predetermined number of revolutions is sent from the controller 204 to the inverters 202 and 203, a driving AC signal or a driving pulse signal having a predetermined frequency (reference frequency) is output from the inverters 202 and 203. It is sent to the motors M ₁ and M ₂ .
Here, when the driving AC signal or driving pulse signal sent from the inverters 202 and 203 deviates from the reference frequency due to the frequency error of the inverters 202 and 203, the rotations of the male and female rotors 101 and 102 are synchronized. Disappear.
However, the driving AC signal or driving pulse signal sent from the inverters 202 and 203 is taken into the feedback circuits 205 and 206, and the frequency error of the inverters 202 and 203 is corrected from the feedback circuits 205 and 206 to match the reference frequency. The control signal to be sent is sent to the inverters 202 and 203.
As a result, the driving AC signal or driving pulse signal sent from the inverters 202 and 203 gradually approaches the reference frequency, and the rotations of the male and female rotors 101 and 102 are synchronized.

  As described above, even if there is a frequency error of the inverters 202 and 203, the feedback circuits 205 and 206 are provided, and the control signal is sent from the feedback circuits 205 and 206 to the inverters 202 and 203 so as to reduce the error. Thus, the rotations of the male and female rotors 101 and 102 are synchronized, the loads on the timing gears 115 and 116 are gradually reduced, and the noise caused by the engagement of the timing gears 115 and 116 can be suppressed. .

  In the above-described embodiment, the twist angle of the screw gear may be continuously changed or may not be changed, or the rotor may be provided with a root portion.

FIGS. 18 and 19 show an advanced embodiment of the vacuum pump shown in FIGS. 14 and 15 described above. The roots 213, 214, screw portions 215 and 216, root portions 217 and 218, screw portions 219 and 220, root portions 221 and 222, and motors M ₁ and M ₂ that are controlled in the same manner as described above are positioned on the left and right. The thing attached to the one side of the rotating shafts 223 and 224 is shown.
By arranging the motors M ₁ and M ₂ in this way, the motors M ₁ and M ₂ can be easily attached to the rotating shafts 223 and 224 even if the diameters of the motors M ₁ and M ₂ are large. 18 and 19, the pair of left and right screws 215, 216, 219, and 220 on the same axis so that the gas sucked from the intake port 225 is divided into left and right and exhausted from the discharge ports 226 and 227 Twisting direction is reversed.
Since other configurations have the same configurations as those in FIGS. 14 and 15, the same reference numerals as those in FIGS. 14 and 15 are used to omit the description thereof.

Next, an embodiment in which a pressure adjusting valve is provided in the vacuum pump according to the present invention will be described with reference to FIGS.
FIG. 20 is a schematic view of the discharge side end face plate portion (inner wall surface portion) of the casing in the screw vacuum pump according to the present embodiment as viewed from the rotor side. FIG. 20 (a) shows the male rotor at the outlet on the male rotor side. It is a figure which shows the state in which a tooth end surface is not located. Further, (b) is a view showing a state in which the male rotor rotates and the tooth end surface of the male rotor is located at the discharge port. FIG. 21 is a diagram schematically showing the screw vacuum pump according to the present embodiment with the circumferential direction of the rotor opened. Further, FIG. 22 is an enlarged view of a main part of the discharge port.

As shown in the figure, reference numeral 301 denotes a male rotor, and a female rotor 302 that meshes with the male rotor 301 is housed in a casing 303 like a conventional screw vacuum pump.
A male rotor end face plate 303a and a female rotor end face plate 303b (FIG. 21) are formed on the discharge side of the casing 303. The end face plate 303a, the end face plate 303b and the tooth end face of the male rotor 301, and the female rotor 302 It is not in contact with the tooth end surface and has a slight interval.
Therefore, the air tightness of the working chambers 301a and 302a is maintained by the male rotor end face plate 303a, the female rotor end face plate 303b, the tooth end face 301b of the male rotor 301, and the tooth end face 302b of the female rotor 302.

Further, discharge ports 304a, 304b, 304c, and 304d are formed in the end surface plate 303a of the male rotor 301. Similarly, discharge ports 305a, 305b, 305c, 305d, and 305e are formed in the end surface plate 303b of the female rotor. ing.
A single discharge port 306 is formed on the end surface plate 303a and the end surface plate 303b so as to straddle the end surface plate 303a and the end surface plate 303b.

Four discharge ports 304 formed in the male rotor side end face plate 303a are provided by one less than the number of teeth of the male rotor (in this embodiment, the number of teeth is five), and four discharge ports 304a to 304d are provided. Are arranged on the pitch circle of the screw gear with the same interval as the tooth pitch of the screw gear constituting the male rotor 301.
Since the discharge ports are formed at the same interval as the tooth pitch of the screw gear constituting the male rotor 301, the male rotor side end face plate 303a can be provided with five discharge ports. The eye discharge port is configured to be a discharge port 306.
Accordingly, the discharge ports 304a to 304d are formed at positions of 72 °, 144 °, 216 °, and 288 ° as shown by the angle from the discharge port 306.

Further, the discharge port 305 formed in the female rotor side end face plate 303b is one from the number of teeth of the female rotor (in this embodiment, the number of teeth is six) like the male rotor side end face plate 303a. A small number of five outlets 305 a to 305 e are arranged on the pitch circle of the screw gear with the same interval as the tooth pitch of the screw gear constituting the female rotor 302.
Since the discharge ports are formed at the same interval as the tooth pitch of the screw gears constituting the female rotor 302, the female rotor side end face plate 303b can be provided with six discharge ports. The eye discharge port is configured to be a discharge port 306.
Accordingly, the discharge ports 305a to 305e are formed at positions of 60 °, 120 °, 180 °, 240 °, and 300 °, as shown by the angle from the discharge port 306.

Since the discharge ports 304a to 304d and the discharge ports 305a to 305e are formed with the positional relationship as described above, the end surface 301b of the screw gear of the male rotor 301 is connected to the discharge ports 304a to 304a as shown in FIG. When 304d is not closed (the end surface 302b of the screw gear of the female rotor 2 closes the discharge ports 305a to 305e), the inside of the discharge ports 304a to 304d is open, and the discharge ports 305a to 305e are open. It is occluded.
Then, when the rotor rotates, the screw gear end face 301b of the male rotor 301 closes the discharge ports 304a to 304d as shown in FIG. 20B from the above state (end face of the screw gear of the female rotor 2). 302b shifts to a state where the discharge ports 305a to 305e are not closed), but the working chamber does not communicate via the discharge ports 304 and 305 in any state.

Next, a discharge valve disposed outside the discharge port will be described with reference to FIG.
Since the basic structure of this discharge valve is the same as that of the conventional discharge valve, the same members as those shown in FIG.
In the figure, a pressure adjusting device 307 is formed integrally with a valve rod 53 that opens and closes the discharge port, a protruding portion 53 a that is formed on the opposite surface of the valve rod 53 and is inserted into the discharge ports 304 and 305, and the valve rod 53. Are formed in the valve box 55 and discharged from the discharge ports 304 and 305. The spring 54 biases the discharge ports 304 and 305 in the closing direction, the valve box 55 that houses the valve rod 53 and the spring 54, and the valve box 55. It is comprised from the air | atmosphere opening | mouth 56 which discharge | releases the performed gas outside.
Here, when the screw pump is arranged vertically and the discharge port 306 is positioned downward, the valve rod 53 is discharged from the discharge ports 304 and 305 when the pressure in the working chamber rises above the atmospheric pressure. Is adjusted to an urging force that supports the dead weight of the valve stem 53.
Therefore, when the pump is placed horizontally, the pressure in the working chamber is higher than the sum of the atmospheric pressure and the urging force of the spring 54 (this value can be said to be almost atmospheric pressure because the urging force of the spring 54 is small). When this happens, the outlets 304 and 305 are opened.

The operation and effect of the embodiment configured as described above will be described in the case where the discharge port is positioned downward.
First, when the pressure of the suction gas is low and the pressure of the working chamber 301a is lower than the atmospheric pressure, the valve rod 53 in the valve box 55 is urged by the spring 54 and closes the discharge ports 304 and 305. At this time, since the protruding portion 53a is inserted into the discharge ports 304 and 305, only a small gap is formed in the discharge ports 304 and 305. Therefore, when the working chambers 301a and 302a are located at and communicated with the discharge ports 304 and 305, the pressure in the working chambers 301a and 302a is not affected by the pressure of the gap between the discharge ports 304 and 305.
Therefore, the gas sucked from the suction port enters the working chambers 301a and 302a formed by the male rotor 301, the female rotor 302, and the casing 303, is compressed by the rotation of both rotors, and is externally discharged from the pressure adjusting device. The ink is discharged from the discharge port 306 without being discharged.
At this time, as shown in FIG. 22, since the inside of the discharge ports 304 and 305 is configured to be closed by the tooth end surface 301b or the tooth end surface 302b of the screw gear constituting the rotor, the working chamber is separated from the adjacent working chamber. There is no communication, and gas does not leak from the high-pressure working chamber to the low-pressure working chamber, and no extra time is required until the intake side reaches a predetermined degree of vacuum.

On the other hand, when the pressure of the suction gas is high and the pressure of the working chamber is higher than the atmospheric pressure, the valve rod 53 is pushed downward, and the gas in the working chamber 51 is put into the valve box 55 from the discharge ports 304 and 305. It is discharged to the outside through the air opening 56 through the clearance.
When the suction pressure decreases and the working chamber pressure does not reach atmospheric pressure immediately before the working chamber communicates with the discharge port, the discharge ports 304 and 305 of the pressure regulator are all blocked, The gas is discharged from the discharge port 306 without being discharged from the pressure adjusting device 307 to the outside.

As described above, in the screw vacuum pump according to the present embodiment, the rotor tooth end surface is located at the discharge port in accordance with the rotation of the rotor of the screw vacuum pump, and the rotor tooth end surface covers the inside of the discharge port. Since it is configured to be closed, gas does not leak from one working chamber adjacent to one working chamber to another working chamber under low pressure without communicating with one working chamber adjacent to the other working chamber. Therefore, no extra time is required until the intake side reaches a predetermined degree of vacuum.
In addition, since the pressure in the working chamber can always be suppressed to almost the atmospheric pressure or less, even when the operation is performed in the state where the suction pressure is close to the atmospheric pressure, the compression is not performed more than necessary. Power consumption can be reduced.
In addition, since the compression is not performed more than necessary, the temperature of the screw vacuum pump does not rise abnormally, and the dimensional accuracy between the engagement between the casing and the rotor, the male rotor and the female rotor, etc., should be improved. it can.

In the above embodiment, the case where four or five discharge ports are provided has been described. However, the present invention is not particularly limited to this, and it is appropriately determined in consideration of the use range, performance, etc. of the screw vacuum pump. Selected.
Further, the discharge port is formed at a position corresponding to the pitch circle of the screw gear constituting the rotor, but is not particularly limited to this position, and may be formed at a position that can be closed by the tooth end face of the screw gear.
Furthermore, in the above embodiment, the spring biasing force has been described as supporting the weight of the valve rod 53. However, the present invention is not limited to this, and the use range, performance, etc. of the screw vacuum pump are taken into consideration. The biasing force of the upper spring may be changed.
Also in this embodiment, the torsion angle of the screw gear may be changed continuously or may not be changed. Also, as shown in FIGS. A root portion (the discharge side end face corresponding to the above-described tooth end face) may be provided on the discharge side.

FIG. 1 shows a conventional screw vacuum pump and is a cross-sectional view taken along the line BB of FIG. FIG. 2 shows a conventional screw vacuum pump and is a cross-sectional view taken along the line AA of FIG. FIG. 3 is a schematic view in which a state in which a male rotor and a female rotor of a conventional screw vacuum pump are engaged is developed in the circumferential direction of the rotor. FIG. 4 is a sectional view of a conventional screw vacuum pump. FIG. 5 is a cross-sectional view of an essential part showing the pressure adjusting device shown in FIG. FIG. 6 is a plan view of a screw gear used in the present invention. FIG. 7 is a development view of the meshing pitch cylinder of the screw gear used in the present invention. The horizontal axis represents the male rolling circumference of the meshing pitch cylinder, the vertical axis represents the amount of progress of the twist, and a parabola (2 FIG. 3 is a development view showing a tooth rolling curve composed of a next curve). FIG. 8 is a diagram showing an increase in the temperature of the screw vacuum pump. The dotted line shows the case of the conventional screw vacuum pump, and the solid line shows the case of one embodiment of the present invention. FIG. 9 is a perspective view of a male and female rotor used in one embodiment of the present invention. 10 is a plan view of the male and female rotors of FIG. 11 is a cross-sectional view of a slew vacuum pump using the male and female rotors shown in FIGS. 12 is a cross-sectional view taken along the line AA in FIG. FIG. 13 is a diagram showing the characteristics of the exhaust speed. FIG. 14 is a cross-sectional view of a screw vacuum pump according to an embodiment of the present invention. 15 is a cross-sectional view taken along line AA in FIG. FIG. 16 is a circuit diagram for controlling the rotation of the male and female rotors shown in FIGS. FIG. 17 is another circuit diagram for controlling the rotation of the male and female rotors. FIG. 18 is a cross-sectional view of a screw vacuum pump according to an embodiment of the present invention. 19 is a cross-sectional view taken along the line AA in FIG. FIG. 20 is a schematic view of a screw vacuum pump according to an embodiment of the present invention as viewed from the discharge side of the casing. FIG. 21 is a view schematically showing a screw vacuum pump according to an embodiment of the present invention with the circumferential direction of the rotor opened. FIG. 22 is an enlarged view of a main part of the discharge port.

Explanation of symbols

1 Male gear (male rotor)
2 Female gear (female rotor)
DESCRIPTION OF SYMBOLS 3 Male mesh pitch cylinder 4 Female mesh pitch cylinder 5 Male tooth shape 6 Female tooth shape 7 Male rotation shaft 8 Female rotation shaft 101 Male rotor 101a Screw gear portion 102 Female rotor 102a Screw gear portion 103 Root portion 104 Root portion 105 Roots Part 106 Roots 109 Casing 109a Suction port (inlet)
109b Discharge port (outlet)
202 inverter 203 inverters 204 controller 205 feedback circuit 206. Feedback circuit M ₁ motor M ₂ motor 301 Oro - motor 301a operating chamber 301b tooth end surface 302 female rotor 302a working chamber 302b tooth end face 303 casing 303a male rotor side end face plate 303b female rotor side End plates 304a to 304d Discharge port 305a to 305e Discharge port 306 Discharge port 307 Discharge valve

Claims (10)

A male rotor and a female rotor that mesh with each other, a casing that houses both rotors, a fluid working chamber formed by the male and female rotors and the casing, and one end and the other end of the working chamber can communicate with each other. In a screw fluid machine comprising a fluid inlet and outlet provided in a casing,
A screw fluid machine characterized in that a twist angle of a screw gear constituting the male and female rotors continuously changes as the twist progresses.   The screw fluid machine according to claim 1, wherein the fluid working chamber formed by the male and female rotors and the casing has suction, continuous compression transfer, and discharge operations.   In the development of the tooth rolling curve on the pitch cylinder diameter of the gear of the male and female rotors, the rolling circumference of the pitch cylinder with respect to the torsional traveling direction is substantially represented by a monotonically increasing function. The screw fluid machine according to claim 1.   4. The screw fluid machine according to claim 3, wherein each screw gear is a gear whose pitch in the rotation axis direction of the screw thread changes, but whose pitch perpendicular to the axis of the screw thread does not change.   A screw gear, characterized in that, in a development view of a tooth rolling curve on a pitch cylinder diameter of a screw gear, a rolling circumference of the pitch cylinder with respect to a direction of the torsion is substantially represented by a monotonically increasing function. A male rotor and a female rotor that mesh with each other, a casing that houses both rotors, a fluid working chamber formed by the male and female rotors and the casing, and a casing that can communicate with one end and the other end of the working chamber. In a screw fluid machine with a fluid inlet and outlet provided,
A screw fluid machine, wherein a root portion is formed at one end of at least one of a screw gear portion and each screw gear portion in the male and female rotors.   The screw fluid machine according to claim 6, wherein a root portion, a screw gear portion, and a root portion are formed in the male and female rotors from the inlet side toward the outlet side. A male screw rotor and a female screw rotor that mesh with each other, a casing that houses both rotors, a fluid working chamber formed by the male and female rotors and the casing, and a casing that can communicate with one end and the other end of the working chamber. In a screw fluid machine with a fluid inlet and outlet provided,
A motor provided in each rotor for driving the male rotor and the female rotor, an inverter for sending a driving AC signal or a driving pulse signal to the respective motors, and a control signal for controlling the frequency of the inverter. A controller for sending out,
A screw fluid machine that controls the number of rotations of the male rotor and the female rotor. A working chamber is formed by a male screw rotor and a female screw rotor that mesh with each other, and a casing that houses both rotors, and the suction gas confined in the working chamber is discharged from the discharge port as the rotor rotates, In the screw fluid machine provided with the pressure adjusting device for limiting the pressure so that the pressure in the working chamber does not increase from the vicinity of the atmospheric pressure,
The pressure adjusting device includes a discharge port formed in a screw end face plate constituting a part of the casing, and a discharge that is provided outside the discharge port and is opened when the pressure in the working chamber exceeds the vicinity of atmospheric pressure. A valve and a tooth end face of the rotor that opens and closes the inside of the discharge port,
A screw fluid machine characterized in that the tooth end surface of the rotor closes the inside of the discharge port in a state where the tooth end surface of the rotor is positioned at the discharge port in accordance with the rotation of the rotor.   The screw fluid machine according to claim 9, wherein the number of the pressure adjusting devices is one less than the number of teeth of the rotor.

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