JPH05195972A - Screw fluid machine - Google Patents

Abstract

PURPOSE:To attain energy saving by setting moment of inertia around the rotary shafts of a male and a female rotors, rotor load torque and static angular transmission error between both rotors in such a way as to satisfy a specific relationship including the number of teeth and the angular frequency of a mesh cycle at the lightest load time. CONSTITUTION:Moment of inertia IM, IF around the respective rotary shafts of a male rotor 1 and a female rotor 1, the gas torque TGF of the female rotor 6 obtained from gas pressure fluctuation based on the geometry of the pressure receiving faces and groove volume change of the rotors 1, 6, the gas torque fluctuation component TGFA of the female rotor 6, the gas torque fluctuation component TGMA of the male rotor 1, and the gear number ratio (r) of the male-female rotors 1, 6 are set in such a way as to satisfy an expression I. In the expression, C1 represents the angle of rotation on the female rotor 6 side so as to represent the harmonic component amplitude, corresponding to the mesh cycle, of static angle transmission error of the female rotor 6 in relation to the male rotor 1. Or omega represents angular frequency corresponding to the mesh cycle, IM and IF represent moment of inertia around the rotary shafts at rotating parts on the male and female rotor sides, and IUF represents equivalent moment of inertia substituted for IM and IF composed onto the female rotor side.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a screw fluid machine for handling air, refrigerant and other gases, and more particularly to rotor and rotary system specifications suitable for obtaining quiet and low-vibration operation at light load.

[0002]

2. Description of the Related Art For screw fluid machines that handle gas, there are an oil supply type that injects oil into the operating space to operate in a gas-liquid mixed phase state, and an operation in the gas phase that does not inject oil into the operating space at all. There is an unlubricated type.

The oil supply system of the present invention has high performance even at low speed operation due to the cooling action, the sealing action or the lubricating action of oil, and is suitable for general-purpose air compressors, air conditioner compressors and vacuum pumps. Widely used.

A refueling type screw fluid machine is proposed, for example, in Japanese Patent Publication No. 40-12437, and the detailed structure is, for example, Japanese Patent Publication No. 42-72 or Japanese Patent Publication No. 42-1002.
No. 7 publication.

As seen in the above-mentioned known example, the screw rotor comprises a pair of male and female rotors having teeth twisted in opposite directions. Examples of tooth profiles used for these rotors are disclosed in Japanese Examined Patent Publication Nos. Sho 56-17559 and Sho 61-8241. Most or all of the tooth profile of the male rotor is formed outside the pitch circle, and conversely, most or all of the tooth profile of the female rotor side is formed inside the pitch circle.

[0006] In a refueling type screw fluid machine, normally, both tooth forms are directly engaged with each other to transmit rotation. Then, normally, the male rotor is on the drive side, and the female rotor is driven from the male rotor via the tooth surface.

The load torque acting on the rotor due to the gas pressure (hereinafter simply referred to as gas torque) is generally mostly applied to the male rotor side and small on the female rotor side. Therefore, particularly when the load is light, the transmission torque between the tooth flanks becomes very small, the meshing becomes unstable, and abnormal vibration and noise accompanied by tooth flank separation are likely to occur.

To prevent this, Japanese Patent Publication No. 56-17559 discloses a tooth profile in which the addendum of the female rotor is large so that the gas torque of the female rotor does not become negative. Further, Japanese Utility Model Laid-Open No. 1-95586 discloses an invention in which an impeller is provided at a shaft end portion of a female rotor to apply a positive torque to prevent tooth surface separation.

[0009]

Conventionally, it has been proposed only to increase the tooth surface transmission torque, but the quantitative limit has not been clarified. There was a problem that abnormal noise was generated when the load was light, depending on the driving conditions. As a countermeasure, the reduction of the load was restricted so that it would not become excessively light when no load is applied.

However, recently, especially in compressors,
In order to save energy, there is an increasing demand for lowering the pressure at the inlet and the outlet when there is no load, and for reducing the mechanical loss even when there is a load. In order to respond to this, the tooth flank transmission torque must be made smaller and smaller, and the possibility of tooth flank separation vibration is further increasing. On the contrary, in order to prevent the tooth surface separation vibration, power saving under no load has been restricted so far.

As the demand for energy saving becomes more severe as described above, it is no longer possible to take a empirical measure by simply limiting the load so that the gas torque of the female rotor does not become too small as in the conventional case. It is necessary to quantitatively capture the tooth flank separation vibration and design the tooth profile, the angle transmission accuracy, and the specifications of the rotating system that are necessary to suppress it.

It is an object of the present invention to provide a screw fluid machine which prevents tooth surface separation vibration and is quiet and which can further reduce power consumption under no load.

[0013]

This is obtained from the moment of inertia of the rotating portion of the male and female rotors about the axis of rotation and the gas pressure calculated based on the geometrical shape of the pressure receiving surface of the rotor and the volume change of the groove. The relation between the load torque of the rotor and the static angle transmission error between both rotors is the lightest load condition among the set operating conditions.

[0014]

[Equation 2]

The rotor rotation system is configured so that

Here, T _GF : Gas torque of female rotor T _GFA : Fluctuation component of gas torque of female rotor T _GMA : Fluctuation component of gas torque of male rotor r: Ratio of male rotor teeth number to female rotor teeth number I _M and I _F : Moment of inertia about the rotation axis of the rotating parts on the male rotor side and the female rotor side, respectively I _0F : Equivalent moment of inertia obtained by combining and replacing I _M and I _F on the female rotor side c ₁ : Rotation angle on the female rotor side The harmonic component amplitude ω of the static angular transmission error of the female rotor with respect to the male rotor, which corresponds to the meshing period, expressed by ω: Angular frequency corresponding to the meshing period

[0017]

With the above construction, the tooth flank transmission torque between the male rotor and the female rotor is always positive even when the load is light,
Tooth surface separation vibration is suppressed, and a quiet and highly reliable screw fluid machine can be obtained.

[0018]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A description will be given below of an oil refueling type screw compressor which is an embodiment of the present invention.

FIG. 1 is a vertical sectional view of the main body of the compressor of this embodiment. The male rotor 1 is supported by bearings 2 and 3 and is rotatably mounted in a casing assembly composed of a main casing 4 and a discharge side casing 5. Similarly, the female rotor 6 is also supported by bearings 7 and 8 and is rotatably mounted in both casings. The main casing 4 and the discharge-side casing 5 are fixed to each other by bolts (not shown). Bearing 3 above
And 8 have both functions of a radial bearing and a thrust bearing.

The male rotor and the female rotor are respectively shown in FIG.
As shown in FIG. 3, teeth 9 and teeth 10 are carved, and a working space is formed by the main casing 4 that covers the rotors and the grooves between the teeth. When both rotors rotate in the directions of arrows 11 and 12, respectively, the volumes of the twisted working spaces change, and the positive displacement compressor operates.
In FIG. 1, a pulley 14 is attached to one end of the shaft 13 of the male rotor.
Is attached and is driven by an electric motor or the like via a belt (not shown). The female rotor is driven directly by meshing with the male rotor. Air enters the suction chamber 15 through an inlet not shown in FIG. 1 and is sucked into the rotor,
After being compressed, it is discharged from the discharge port 16 to the outside of the machine.

A shaft seal device 17 is attached to the drive shaft 13, and seals the outer peripheral surface of the shaft 13 so that oil lubricating the bearings 2 and 7 does not leak to the outside of the compressor. A hole 18 is provided in the side wall of the main casing 4 to inject oil from the outside toward the rotor. This oil not only cools the working air and seals the clearance of the sliding portion, but also lubricates the meshing portion and is discharged from the discharge port 16 together with the compressed air.

Oil is also supplied from another oil supply port 19, lubricates the suction side bearings 2 and 7, and then is sucked into the rotor through the clearances 20 and 21 between the shaft hole and the shaft. further,
Oil is also supplied from another oil supply port 22, part of which directly enters the bore for housing the rotor, and the other part lubricates the discharge side bearings 3 and 8 and then flows into the bore through the hole 23. These oils mix with the above-mentioned jet oil in the rotor and are discharged to the discharge port together with the gas.

The inner wall of the bore facing the rotor groove during the suction stroke is provided with relief portions 24 and 25 in FIGS. 1 and 2. This is to reduce oil-mediated sliding loss between the addendum and the inner wall of the bore.

FIG. 3 shows a system diagram of an apparatus for operating the compressor of this embodiment. The compressor main body 26 is the same as that shown in FIGS. 1 and 2, and a schematic cross-sectional view taken along the section line II-II of FIG. 1 is shown and details thereof are omitted.

Air is sucked into the capacity adjusting valve 27 from the atmosphere as shown by an arrow 28 and enters the suction chamber 15 of the compressor through the valve 29. The air compressed by the rotor is discharged into the discharge chamber 30, and enters the oil separator 32 via the pipe line 31. Here, the air and the oil are separated, and the air is pressure-fed to the air use line through the conduit 33 and the check valve 34 in the middle thereof.

Oil collects in the oil tank 35 below the oil separator,
It is pushed out by the pressure of the discharge air, passes through the oil cooler 36 and the oil strainer 37, and is again supplied to the compressor body 26. In this figure, the oil supply path is represented by the injection hole 18, and the other supply paths are omitted.

When the amount of air used decreases, the air pressure in the conduit 33 rises, and the controller 38 detects this and sends a signal to the capacity adjusting valve 27 to decrease the opening of the valve 29. As a result, the air pressure in the suction chamber 15 is reduced, the power of the compressor is reduced, and the discharge air amount is reduced.

The opening of the valve 29 changes depending on the amount of air used, and the suction pressure also changes correspondingly. Normally, the suction pressure is reduced to 13 kPa. In the compressor of this embodiment, after reducing the suction pressure to 13 kPa, a command is issued from the controller 38 to further reduce the power, and the pipeline 33 is sent.
The solenoid valve 39 connected to is opened. As a result, high-pressure air is released into the atmosphere, the discharge pressure of the compressor is lowered, and the power is reduced.

The torque calculation method of the screw compressor of this embodiment will be described below. The pressure change required for torque calculation is assumed as follows.

In FIG. 4, the horizontal axis shows the working space volume, and the vertical axis shows the pressure of the air therein. V _S and V _D respectively represent the working space volume when the working space is closed from the suction port and when it approaches the discharge port. V _S
/ V _D is commonly referred to as the built-in volume ratio (v _i ).
P _S and P _D are suction pressure and discharge pressure, respectively.

Since the volume on the horizontal axis is a geometrical quantity, if the shape of the rotor is determined, it is uniquely determined as a function of the rotor rotation angle regardless of the calculation method.

The change at full load proceeds as A ₁ -A ₂ -A ₃ -A ₄ . Here, between A ₁ -A ₂ is the intake stroke, while A ₂ -A ₃ is the compression stroke, A ₃ -A ₄ while is discharge stroke. It is assumed that compression progresses due to isentropic change between A _{2 and} A ₃ , and when the assumption of an ideal gas is approximately satisfied, the relationship between P and V is calculated by the following equation.

[0033]

[Equation 3]

Where κ is the specific heat ratio of the working gas. Since air is also regarded as an almost ideal gas, κ is set to 1.4 and calculated by equation (2).

The discharge port opens from the position A ₃ ,
The position of the discharge port is determined so that the pressure in the working space is equal to P _D when the change is approaching A ₃ .

When the suction pressure is reduced to the minimum value set in the compressor, the change proceeds as A ₅ -A ₆ -A ₇ -A ₃ -A ₄ . Although the discharge port opens at the position of A ₇ , the pressure in the working space at this time is considerably lower than P _D , and air flows backward from the discharge chamber into the working space along with the port opening, and the pressure in the working space rises to P _D. To do.

To further suppress the load power of the compressor,
In this embodiment, the discharge pressure is reduced while the suction pressure is reduced as described above. The discharge pressure at this time is shown by P _DU in the figure.
Change proceeds as _{A 5 -A 6 -A 7 -A 8} -A 9, the load torque in this state smallest.

In this example, the pressure change is calculated by assuming an ideal change without internal leakage or flow loss. However, when it is necessary to perform a strict evaluation, the internal leakage and flow loss are considered by simulation. The calculated pressure change may be calculated, or the actual pressure change may be measured and used.

The load torque T acting on the rotor can be calculated by the following equation for one working space. That is,

[0040]

[Equation 4]

Where P is the pressure, M is the area moment of inertia about the rotation axis of the rotor surface forming the working space, and M is a geometrical quantity like the volume. Therefore, if the shape of the rotor is determined, the calculation method is calculated. It is uniquely determined regardless of.

For example, the first moment of area can be calculated as follows. As shown in FIG. 5, the rotor is placed in a three-dimensional Cartesian coordinate system (x, y, z). Here, the z-axis is the rotation axis direction of the rotor, and is positive from the suction side toward the discharge side. The y-axis is 90 degrees from the x-axis in the normal rotation direction of the rotor.
The position is rotated once. One working space 10 on the rotor
The rotor surface forming 1 is projected perpendicularly to the plane X and the plane Y perpendicular to the plane X and the plane Y, respectively.

The projected figures are as shown by 102 and 103 in the figure, and an example of their contours is shown in FIG. However, this figure is generally shown and does not necessarily represent a faithful shape.

The area moment is obtained by integrating the following quantities on the plane X and the plane Y along the outline of the projected figure. That is, for plane X,

[0045]

[Equation 5]

For plane Y,

[0047]

[Equation 6]

The area moment M is

[0049]

[Equation 7]

Since the rotor is engraved with a plurality of grooves,
The gas torque of the rotor is the sum of the above torques calculated for the working space of each groove.

Now, it is assumed that the tooth surfaces of the male rotor and the female rotor are engaged with each other and rotate without being separated from each other.
It is assumed that a gas torque that changes with rotation acts on both rotors, and that the engagement of both rotors has an angle transmission error that changes in synchronization with the engagement.

If both rotors, including the shaft and the pulley, are regarded as a rigid body, and the spring force of the belt is weak, it is assumed that the rotating system beyond the belt is vibrationally insulated. Equations (7) and (8) are as follows. That is,

[0053]

[Equation 8]

[0054]

[Equation 9]

Here, θ is the rotor rotation angle, and the normal rotation direction of each rotor is positive.

T: time I: moment of inertia about the rotation axis T _G : gas torque acting on the rotor, the reverse direction of the rotation direction being positive.

T _FM : Tooth surface transmission torque received by the male rotor from the female rotor T _MF : Tooth surface transmission torque received by the female rotor from the male rotor T _S : Drive torque transmitted from the belt to the male rotor, which is a constant value.

The subscripts M and F represent quantities relating to the male rotor and the female rotor, respectively.

The following equations (9) and (10) are established under the condition that the rotor is not separated.

[0060]

[Equation 10]

[0061]

[Equation 11]

Where r is the ratio of the number of male rotor teeth to the number of female rotor teeth, δ is the angular transmission error of the female rotor rotation angle with respect to the male rotor. Here, the gas torque is divided into a constant term and a variation term, and equation (1)
It is expressed as 1) and (12).

[0063]

[Equation 12]

[0064]

[Equation 13]

The suffixes 0 and A at the end represent a constant term and a variation term, respectively.

From its definition, T _S is the sum of the male rotor gas torque average value and the torque obtained by converting the female rotor gas torque average value to the male rotor side,

[0067]

[Equation 14]

It becomes

From the above equations, the tooth surface transmission torque T _MF received by the female rotor from the male rotor is calculated as follows:

[0070]

[Equation 15]

It becomes Here, I _0F is an equivalent moment of inertia in which I _M and I _F are both replaced by the female side, and is expressed by the formula (15).
become that way.

[0072]

[Equation 16]

In T _MF, the normal rotation direction of the rotor is positive. T _MF must always be positive in order for flank separation not to occur. That is,

[0074]

[Equation 17]

The angle transmission error δ is a periodic function that varies in synchronization with the tooth surface engagement period, and is represented by the equation (17) in the form of Fourier series.

[0076]

[Equation 18]

Where c _{n is} the half-amplitude value of the angular transmission error of the n-th component of the meshing γ _n is the phase angle of the angular transmission error of the n-th component of the meshing ω is the angular frequency corresponding to the meshing period, and the meshing is If the frequency is f and the pi is π,

[0078]

[Formula 19]

Substituting equation (17) into equation (16),

[0080]

[Equation 20]

If the angle transmission error δ is known in advance and the tooth surface separation is evaluated in detail, the above formula may be used, but this is not always practical in a general production site. Therefore, the procedure is simplified as follows.

First, since the tooth surface separation vibration generally occurs at the meshing frequency, the harmonic component of the angle transmission error is ignored in equation (19), and only the meshing first-order component is considered.

The phase angle γ _n of the angle transmission error can take any value from 0 to 2π, but when the possibility of tooth surface separation is judged by the equation (19), the difference between the phase angle of the gas torque and the phase angle of the angle transmission error is determined. Must assume that it will take every case. Therefore, the second item on the right side of equation (19) is the time t.
We will take the maximum value regardless of.

Considering the above points, the following equation is used for the determination of tooth flank separation.

[0085]

[Equation 21]

On the contrary, when the gas torque and the moment of inertia of the rotating portion are known, the angle transmission accuracy c ₁ required between the male and female tooth flanks is

[0087]

[Equation 22]

Since the angle transmission error differs depending on each rotor or each assembly of the compressor, the allowable error range is set at the design stage and the tooth surface separation is evaluated by the upper limit value.

The angle transmission error is as follows:
It can be measured with a mesh tester. However, more strictly, it is better to measure with the rotor incorporated in the compressor. To do this, attach an angle detection device to the shaft ends of both male and female rotors, measure each rotation angle while slowly rotating both rotors from the drive shaft side, and obtain the relative rotation angle difference. Good.

The tooth profile used in the compressor of this embodiment is shown in FIG. In the figure, 110 and 111 are the centers of the male rotor and the female rotor, respectively, and 112 and 113 are the pitch circles of the male rotor and the female rotor, respectively. Each part of the tooth profile is configured as follows.

Male rotor side 114-115: Centered at 110: 110, pitch circle 112
An arc with a radius equal to.

115-116: Tooth profile 121 of female rotor
A curve created by the arcs that make up 122.

116 to 117: An arc having a center on a line segment connecting 110 and 111 and a radius of R ₂ .

117-118: Female rotor tooth profile 123-
A curve created by the parabola that constitutes 124.

118-119: Tooth profile 124 of female rotor
A curve created by the arcs that make up 126.

Female rotor side 120-121: Centered at 111: 111, pitch circle 113
An arc with a radius equal to.

121-122: An arc having a center on a line segment connecting 111 and 121 and a radius of R ₁ .

122-123: Male rotor tooth profile 116-
The curve created by the arcs that make up 117.

123-124: The vertex is at 123, the focal point is at the point 125 on the line segment connecting 110 and 111,
A parabola whose focal length is a.

124-126: An arc having a center on the line segment connecting 111 and 126 and a radius of R ₃ .

The rotor specifications of this embodiment are shown below.

Male rotor Female rotor Number of teeth 5 6 Rotor outer diameter (mm) 160 128 Rotor length (mm) 224 Center distance (mm) 117 Full winding angle (°) 300 250 Displacement volume (cm ³ / rev) 454 v _i 4.4 The half-amplitude value of the first-order component of the meshing angle transmission error is set to 10 seconds here, and the female rotor torque is calculated below.

The dimensions of each part of the tooth profile are determined as follows using the symbols in FIG. These values were selected by conducting a parameter survey so that the gas torque on the female rotor side would be sufficiently large.

A = 19.2 mm R ₁ = 2.5 mm R ₂ = 10.0 mm R ₃ = 6.3 mm In this specification, the moment of inertia around the rotation axis including the rotor shaft and pulley is the male rotor and the female rotor. is 0.111kg · m ² and 0.015 kg · m ² respectively, for.

The calculated torque of the female rotor of this embodiment is shown in FIGS. 8 to 10 for various operating conditions. Arrows (a) and (b) in each figure show the female rotor gas torque T _GF of the formula (1) and the value on the right side of the formula (1), respectively.

FIG. 8 shows that at full load, that is, the discharge pressure is 934 k.
Pa shows the torque when the suction pressure is atmospheric pressure (101 kPa), and (a) always exceeds (b).

FIG. 9 shows the torque when the suction pressure is reduced to 13 kPa while keeping the discharge pressure at 934 kPa, and similarly (a) always exceeds (b).

FIG. 10 shows the torque when the suction pressure is reduced to 13 kPa and the discharge pressure is reduced to 268 kPa. This is the lightest load condition set for the compressor. Also in this case, (a) always exceeds (b), and the female rotor tooth surface is pressed against the male rotor tooth surface by the gas pressure.

Therefore, quiet running can be performed without tooth surface separation under any set operating conditions.

[0110]

According to the present invention, the tooth flank transmission torque under a light load is always positive even when there is an angle transmission error between the male and female rotors, the tooth surface separation vibration is suppressed, and the noise is quiet and reliable. A high screw fluid machine is obtained. In particular, in order to save power when there is no load, even when selecting pressure conditions that result in a lower torque, the rotational system specifications and the required rotor angle transmission error are determined according to the gas torque, and tooth surface separation vibration Therefore, the power consumption can be further reduced as compared with the conventional one.

[Brief description of drawings]

FIG. 1 is a cross-sectional view of a screw compressor according to an embodiment of the present invention.

FIG. 2 is a vertical cross-sectional view taken along the section line I-I of the compressor shown in FIG.

FIG. 3 is a system diagram of an air compression device using the compressor shown in FIGS. 1 and 2.

FIG. 4 is an explanatory view of volume-pressure change of the compressor.

FIG. 5 is an explanatory diagram of a projection method of a working space.

FIG. 6 is an explanatory diagram of an area first-order moment calculation method.

FIG. 7 is an explanatory diagram of a rotor tooth profile used in the compressor of the present invention.

FIG. 8 is a female rotor torque characteristic diagram at full load of the compressor of the present invention.

FIG. 9 is a female rotor torque characteristic diagram when the intake air of the compressor of the present invention is blocked.

FIG. 10 is a female rotor torque characteristic diagram of the compressor of the present invention when intake air is blocked and discharge pressure is reduced.

[Explanation of symbols]

1 ... Male rotor, 2, 3, 7, 8 ... Bearing, 4 ... Main casing, 6 ... Female rotor, 14 ... Pulley, 15 ... Suction chamber,
16 ... Discharge port, 27 ... Volume adjustment valve, 34 ... Check valve, 38
… Controller, 39… Solenoid valve.

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Fumio Takeda Futao Takeda 502 Jinritsucho, Tsuchiura-shi, Ibaraki Hiritsu Manufacturing Co., Ltd.Mechanical Research Institute (72) Inventor Naoyuki Tanaka 502 Jinritsu-cho, Tsuchiura-shi, Ibaraki Hiritsu Manufacturing Co., Ltd. Inside the Mechanical Research Laboratory (72) Inventor Hirotaka Kamiya 502 Kintatecho, Tsuchiura City, Ibaraki Prefecture

Claims (1)

[Claims] 1. A pair of male and female rotors having teeth twisted in opposite directions, a pair of bores intersecting each other and containing the male and female rotors, a gas inlet and a gas outlet. , A casing for rotatably instructing the male rotor and the female rotor via bearings, and the teeth of the male rotor are formed such that main portions or all of them are formed outside a pitch circle,
The main part or all of the teeth of the female rotor are formed inside the pitch circle, the shaft directly connected to the male rotor is driven by an external prime mover via a belt and a pulley, and the female rotor is connected to the male rotor. In the screw fluid machine driven by the meshing of the rotors, the moment of inertia about the rotation axis of the rotating portions of the female and male rotors, the geometric shape of the pressure receiving surface of the rotor, and the gas pressure fluctuation based on the volume change of the groove, The relation between the load torque of the rotor and the static angle transmission error between the two rotors obtained from The screw fluid machine is characterized in that the rotor rotation system is configured so that Here, T _GF : Gas torque of the female rotor T _GFA : Fluctuation component of the gas torque of the female rotor T _GMA : Fluctuation component of the gas torque of the male rotor r: Ratio of the number of teeth of the male rotor to the number of teeth of the female rotor I _M and I _F : Inertia moments around the rotation axis of the rotating portions on the male rotor side and the female rotor side, respectively, I _0F : Inertia moments obtained by equivalently synthesizing and replacing I _M and I _F on the female rotor side c ₁ : Harmonic component amplitude corresponding to the meshing period of the static angle transmission error of the female rotor with respect to the male rotor, represented by the rotation angle on the female rotor side ω: Angular frequency corresponding to the meshing period