JPH08152047A - Balancing device for rotating machine - Google Patents

Abstract

PURPOSE: To balance the rotor of a rotating machine according to the change of a vibration state without stopping operation of the rotating machine. CONSTITUTION: The balancing device of a rotary machine comprises one set of balancing weight moving pipes 12a and 12b arranged at the outer periphery of a rotor 1: and balancing weights 15a and 15b contained in the respective balancing weight moving pipes 12a and 12b. The balancing device has a hydraulic control device 7 located at the outside of the rotor 1. The balancing weights 15a and 15b are movable peripherally of the rotor 1 by means of an oil pressure in the balancing weight moving pipes 12a and 12b. The oil pressure control device 7 controls an oil pressure in the balancing weight moving pipes 12a and 12b to move the balancing weights 15a and 15b to an arbitrary position in the peripheral direction of the rotor 1.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a balancing apparatus for balancing a rotor of a rotary machine such as a steam turbine to reduce vibration of the rotary machine, and more particularly to a change in vibration state during operation of the rotary machine. BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a rotary machine balancing device capable of correspondingly balancing rotors.

[0002]

2. Description of the Related Art Conventionally, as a rotary machine such as a steam / gas turbine or a generator used in a plant for thermal power / nuclear power generation, a so-called multi-shaft / bearing system configuration as shown in FIG. What is taken is known. In FIG. 8, a high-intermediate-pressure rotor 31, a low-pressure rotor 32, and a generator rotor 33 that form a rotary machine each have a rotor shaft 30, and the rotor 31 and the rotor 32, and the rotor 32 and the rotor 33 include the rotor shaft 30. They are directly connected by a coupling 2 that connects the two. In addition, each rotor 31, 3
The two and 33 rotor shafts 30 are supported at both ends by journal bearings 34.

Usually, the rotor 31 of such a rotary machine,
Since the weight of 32, 33 reaches from several tons to several tens of tons,
When vibration occurs in the rotating machine due to the rotation, the energy of the vibration becomes extremely large. Therefore, in order to prevent the vibration of the rotating machine, each rotor 3 is shipped from the factory.
High-precision balancing is performed so that unbalance does not remain in 1, 32, and 33. Furthermore, during on-site assembly of the rotating machine, unnecessary force acts on the rotors 31, 32, 33 due to misalignment of the coupling 2, which couples the rotors 31, 32, 33, surface opening, etc. It is considered that the axial center positions of the rotors 31, 32, 33 are adjusted by the respective bearings 34 so that the rotors 31, 32, 33 are in the optimum alignment state.

However, a rotary machine using a high-temperature working fluid such as a steam / gas turbine is subject to thermal deformation or alignment change of the rotor due to temperature change of the working fluid.
The vibration state may change significantly. FIG. 9 shows a typical DSS as an example of such a vibration change of a rotary machine.
It is the figure which showed the amplitude change of the vibration at the time of the start of the high-pressure rotor bearing part of the large-sized thermal power generation turbine used for (daily starting and stopping) and WSS (weekly starting and stopping). In FIG. 9, Ho
The t state represents a state in which the operation was stopped the night before and started the next morning, and the Cold state represents a state in which the operation was restarted after a long-term operation stop. As shown in FIG.
In the t state, the balance of the rotor 1 is almost perfect,
Almost no vibration is generated, but in the cold state, the rotor 1 is out of balance, and vibration having a considerably larger amplitude than in the hot state is generated. Furthermore, Col
In the d state, the vibration state changes as the rotation speed increases, and the amplitude fluctuates greatly.

FIG. 10 is a diagram showing, as another example of the vibration change of the rotary machine, the vibration change of the high-pressure rotor bearing portion due to the load change of the steam turbine. In FIG. 10, the radial direction coordinates indicate the vibration amplitude, and the circumferential direction coordinates indicate the vibration phase (counterclockwise). As shown by the solid line in FIG. 10, the steam turbine reaches the rated speed and the load is 0M.
The amplitude is about 0.02 mm when W, but the vibration state changes as the load rises, and the amplitude increases to about 0.05 mm when the load is 500 MW.

Furthermore, in recent years, the bearing spacing of rotary machines has increased due to the lengthening of rotary machines such as steam turbines, and the weight of rotors has also increased. Therefore, the natural frequency of the rotor system of such rotary machines is generally It tends to decrease and become a complicated vibration system in which many natural vibration modes exist below the rated speed. Therefore, the sensitivity of the natural vibration mode of the rotor of the rotating machine is increased, and the temporary bending of the rotor due to the change of the working fluid temperature and pressure of the rotating machine and the vibration change due to the change of the alignment are likely to occur.

[0007]

Conventionally, as shown by a broken line in FIG. 10, for a rotary machine having the above-described vibration change characteristic, the change of the vibration vector is optimum as a whole (the maximum amplitude value is Vibration is reduced by balancing the rotor with a constant vibration state as a reference (to minimize) or by reducing the sensitivity of the natural vibration mode by changing the design of the bearing and the rotor.

However, even if the variation of the vibration vector is optimized by balancing the rotor based on the constant vibration state,
Balancing work of a rotary machine that constitutes a complicated vibration system as described above becomes extremely difficult. In addition, despite such difficult work, it is difficult to achieve sufficient vibration reduction in response to vibration changes over the entire range of operating conditions of the rotating machine. For example, in the case of FIG. 10, the vibration state with the maximum amplitude of about 0.03 mm remains even after the optimization of the change in the vibration vector indicated by the broken line. Further, as shown in FIG. 11, a vibration state in which the amplitude is extremely large during the vibration change (solid line)
If (includes) (solid line), even after the optimization of the change in the vibration vector (dashed line), a vibration state with a very large amplitude still remains, which may hinder the operation of the rotating machine. . Further, even if the sensitivity of the natural vibration mode is reduced by changing the design of the bearing and the rotor, the loss due to the suspension of the operation of the rotating machine for a long period of time and the suspension of the operation of the plant such as the power generation facility due to the suspension will be enormous.

Therefore, as a means for reducing the vibration in response to such a vibration change of the rotating machine during operation, a magnetic damper for applying a magnetic attraction force on the outer circumference of the rotor to damp the vibration change, or the bearing itself. It is also conceivable to use a magnetic bearing that actively reduces vibration by giving a magnetic attraction force to. However, although these means have a track record for small rotating shaft systems, they do not have a track record for large rotating machines such as steam turbines, and are particularly reliable for rotating machines with multi-axis / bearing systems as described above. There is a problem in terms of sex.

The present invention has been made in consideration of such a point, and the rotor of the rotary machine can be balanced according to the change of the vibration state without stopping the operation of the rotary machine. It is an object of the present invention to provide a balancing device.

[0011]

According to the present invention, a ring-shaped counterweight moving pipe provided on the outer circumference of a rotor of a rotary machine and a hydraulic pressure in the counterweight moving pipe housed in the counterweight moving pipe. A counterweight that is movable in the circumferential direction of the rotor, and a hydraulic control device that controls the hydraulic pressure in the counterweight moving pipe to move the counterweight to an arbitrary position in the rotor circumferential direction. It is a balancing device for a rotary machine, characterized by comprising:

[0012]

According to the present invention, the hydraulic pressure in the counterweight moving pipe is controlled by the hydraulic control device to move the counterweight in the weight moving pipe in the circumferential direction of the rotor, and the counterweight is used to fish the rotor of the rotating machine. It can be moved to a predetermined position for alignment.
Therefore, without stopping the operation of the rotating machine, the rotor can be balanced in accordance with the change in the vibration state, and the vibration of the rotating machine can be reduced.

[0013]

Embodiments of the present invention will now be described with reference to the drawings. First, an embodiment of the present invention will be described with reference to FIGS.

In FIG. 1, a balancing device for a rotary machine is provided near a coupling 2 of a rotor (rotor shaft portion) 1. This balancing machine for a rotating machine includes a pair of counterweight moving pipes 12a and 12b provided on the outer periphery of the rotor 1 side by side in the axial direction of the rotor 1, and the counterweight moving pipe 1
Counterweights 15a stored in 2a and 12b,
15b and. The balancing machine for a rotary machine is provided with a hydraulic control device 7 including a hydraulic unit 7 and a control device 8 outside the rotor 1.

Next, the counterweight moving pipes 12a and 12b and the counterweights 15a and 15b will be described in detail with reference to FIGS. Among them, the counterweight moving pipe 12a and the counterweight moving pipe 12b, and the counterweight 15a and the counterweight 15
Since each b has the same configuration, hereinafter, the counterweight moving pipe 12
A and the counterweight 15a will be described. As shown in FIGS. 1 and 2, the counterweight moving pipe 12 a is a ring-shaped pipe having a circular cross section, and the counterweight 15 a is the counterweight moving pipe 1.
Corresponding to the inner space of 2a, the ring has a shape like a ring. Further, as shown in FIG. 3, the counterweight 15a has an outer diameter slightly smaller than the inner diameter of the counterweight moving pipe 12a. Further, seal rings 16 are fitted on the outer circumferences of both ends of the counterweight 15a,
The seal ring 16 slidably and closely contacts the inner wall surface of the counterweight moving pipe 12a.

Further, as shown in FIG. 2, the counterweight moving pipe 1
A partition plate 14 for partitioning the internal space of the counterweight moving pipe 12a into two is provided at a portion of 2a corresponding to the rotor 0 mark (reference point of the circumferential position of the rotor 1) 13.
In FIG. 2, the partition plate 14 inside the counterweight moving pipe 12a
The space from the left side to the end face of the counterweight 15a in the counterclockwise direction (more precisely, up to the seal ring 16 as shown in FIG. 3) is filled with high-pressure oil for hydraulic drive to form the first oil chamber 18. ing. Similarly, the space from the right side of the partition plate 14 inside the counterweight moving pipe 12a to the end surface of the counterweight 15a in the clockwise direction forms the second oil chamber 19.

Next, the hydraulic control device 7 will be described with reference to FIG. As shown in FIG. 1, the hydraulic unit 9 of the hydraulic control device 7 includes two sets of oil pipes 6a ₁ , 6a ₂ , 6b ₁ , 6b _2.
Is linked to Then, the hydraulic unit 9 responds to the control signal from the control device 8 by the oil pipes 6a ₁ , 6a ₂ , 6
The high-pressure oil is supplied or recovered through b ₁ and 6b ₂ , and the pressure of the high-pressure oil supplied or recovered from each oil pipe 6a ₁ , 6a ₂ , 6b ₁ and 6b ₂ can be adjusted independently.

Next, referring to FIG. 1 and FIG. 2, the connecting structure of the counterweight moving pipes 12a and 12b and the hydraulic unit 9 of the hydraulic control device 7 will be described. In addition, the counterweight moving pipe 1
Since the connecting structure of 2b and the hydraulic unit 9 of the hydraulic control device 7 is similar to that of the counterweight moving pipe 12a, first, the counterweight moving pipe 12a and the hydraulic unit 9 of the hydraulic control device 7 are connected. The connection structure will be described in detail.

As shown in FIG. 1, a set of oil coupling devices 3a ₁ and 3a ₂ are arranged at a constant distance in the axial direction from the counterweight moving pipe 12b on the outer periphery of the rotor 1 at a constant distance. Has been done. Each of the oil coupling devices 3a ₁ and 3a ₂ includes a rotor 1
-Shaped oil joint 5 with a U-shaped cross section that extends in the circumferential direction of
And a pair of ring-shaped oil seals 4 interposed between the oil joint 5 and the outer peripheral surface of the rotor 1. The oil coupling devices 3a ₁ and 3a ₂ are stationary even when the rotor 1 is rotating, and the oil seal 4 prevents the oil coupling devices 3a ₁ and 3a ₂ and the rotor 1 from rotating.
It is possible to prevent oil leakage while maintaining a sliding contact with. In addition, the oil groove 10 is formed in each portion of the outer peripheral surface of the rotor 1 corresponding to the oil coupling devices 3a ₁ and 3a _2.
Are formed. The bottom of the oil groove 10a and the first oil chamber 18 of the counterweight moving pipe 12a are connected to each other by the connecting pipe 11a _1.
Communicating a bottom of the oil groove 10b, and a second oil chamber 19 of the balancing weight moving pipe 12a communicates with the connecting pipe 11a ₂ (see FIG. 2).

On the other hand, the oil joints 5 of the respective oil coupling devices 3a ₁ and 3a ₂ and the hydraulic unit 9 of the hydraulic control device 7 are connected by oil pipes 6a ₁ and 6a ₂ , respectively. in this way,
The first oil chamber 18 and the second oil chamber 1 of the counterweight moving pipe 12a
9 is independently connected to the hydraulic unit 9 of the hydraulic control device 7. Then, by controlling the hydraulic unit 9 by the control device 8 of the hydraulic control device 7, the hydraulic pressure P ₁ of the first oil chamber 18 and the hydraulic pressure P ₂ of the second oil chamber 19 of the counterweight moving pipe 12a (see FIG. 3). ) Respectively, and it is possible to supply or discharge / recover high-pressure oil to the first oil chamber 18 and the second oil chamber 19 in accordance with the pressure difference between the oil pressure P ₁ and the oil pressure P _2. Has become.

The connection structure of the counterweight moving pipe 12a and the hydraulic unit 9 of the hydraulic control device 7 has been described above. As shown in FIG. 1, the counterweight moving pipe 12b and the hydraulic unit 9 of the hydraulic control device 7 are described. Also the connecting pipes 11b ₁ and 11
b ₂ , the oil coupling device 3 b ₁ , 3 b ₂ , and the oil pipe 6 b ₁ ,
Similarly connected via 6b ₂ .

A position sensor 20 shown in FIG. 2 has a counterweight 15 based on the rotor 0 mark 13 as a reference.
This is for detecting the positions of a and 15b in the circumferential direction of the rotor 1.

In this embodiment, a balancing device is provided in the vicinity of the coupling 2 of the rotor (rotor shaft portion) 1. However, this is a disk between rotor blades of a steam / gas turbine or the like. This is because there are many space restrictions due to the large number of components and the like, and that the vicinity of the coupling 2 is likely to become an antinode of the vibration mode in the multi-axis / bearing system. Here, the "antinode of the vibration mode" means a place where the amplitude becomes large when a certain natural vibration mode is excited, and the vibration of the rotating machine is effectively reduced by providing a balancing device at this place. Can be expected. Therefore, depending on the type and structure of the rotating machine, the position of the antinode of the vibration mode, and the like, the balancing device may be provided at another place on the rotor.

Next, referring to FIGS. 4 to 6, the counterweight 15 will be described.
The principle of balancing the rotor 1 with a and 15b will be described in detail. Generally, the balance of the rotating body is performed by two-sided balance in which a correction surface perpendicular to the rotation axis is provided at two different points on the rotation axis, and an eccentric weight for correction is added on this surface. Then, in order to correct and balance the unbalance of the rotating body with the sum of these unbalance vectors (weight W (g) x eccentric radius R (mm), angle (°) around the rotation axis) due to the eccentric weight for correction. The correction amount of can be expressed.

In this embodiment, as shown in FIG. 4, the angle around the rotation axis of the counterweights 15a and 15b is set to the rotor 0.
It is set to θ ° counterclockwise with reference to the mark position 13.
For example, when the counterweights 15a and 15b are diametrically opposed to each other as shown by line AA in FIG.
Since the unbalance vectors of 5a and 15b have the same magnitude and opposite directions, the correction amount represented by the sum of the unbalance vectors is 0 gmm regardless of the value of θ. Further, for example, a counterweight 15 having a weight W (g) and a mounting radius R (mm)
When a and 15b are at the positions shown by the line BB in FIG. 4, the correction amount is (2 × W × R × cos 30 ° (gm
m), 200 °). In this way, by changing the positions of the two balancing weights 15a and 15b, (0 gmm
The correction amount can be adjusted in the range of 2WR (gmm), 0 ° to 360 °.

Next, referring to FIGS. 5 and 6, a method of balancing the rotor 1 in a certain vibration state with the balancing weights 15a and 15b will be described. For example, in the vibration vector diagram of FIG. 5, it is assumed that the rotor 1 is in the vibration state represented by the point C1. At this time, the vibration vector E1 indicated by the broken line arrow in FIG.
This is a variation of the vibration vector when unbalanced (10000 gmm, 0 °) is attached to the position corresponding to the counterweight moving tubes 12a and 12b of the rotor 1, and is called an influence coefficient vector.

For example, as shown in FIG. 5, the influence coefficient vector E1 is (0.02 mm / 10000 gmm, 10 °).
And the vibration vector V1 of the vibration state C1 is (0.04
mm, 210 °). In this case, the unbalance vector (correction amount) W1 to be added to reduce the vibration in the vibration state C1 is obtained by the following formula.

V1 + W1 × E1 = 0 ∴W1 = −V1 / E1 By substituting a specific numerical value into this formula, W1 = − (0.04 mm, 210 °) / (0.02 mm / 10000 gmm, 10 °) = -(20000 gmm, 200 °) = (20000 gmm, 200 ° -180 °) = (20000 gmm, 20 °) The method of obtaining the correction amount W1 as described above is generally called the influence coefficient method.

Such an unbalance vector (correction amount) W1
For example, weight W = 100 g, mounting radius R = 200 m
In order to create the counterweights 15a and 15b of m, the counterweights 15a and 15b may be moved to the positions shown in FIG. That is, in FIG. 6, the counterweight 15
The correction amount by a and 15b is (2 × 100 g × 200 mm × cos 60 °, 20 °) = (20000 gmm, 20 °) = W1.

Next, the operation of this embodiment having the above structure will be described.

When the rotor 1 is balanced by the balancing device, first, the predetermined positions in the circumferential direction of the rotor 1 of the balancing weights 15a, 15b for balancing the rotor 1 are obtained by the above-described influence coefficient method. Next, the hydraulic unit 9 is controlled by the control unit 8 of the hydraulic control unit 7 to move the counterweight moving pipe 12
The counterweights 15a and 15b in a and 12b are moved to the predetermined positions. The movement of the counterweights 15a and 15b by the hydraulic control device 7 will be described below in detail for the counterweight 15a.

For example, when the counterweight 15a is moved counterclockwise in FIG. 2, the hydraulic unit 9 is controlled by the control unit 8 of the hydraulic control unit 7, and the oil pipes 6a ₁ and 6a ₂ connected to the hydraulic unit 9 are controlled. Through counterweight moving tube 12a
Oil pressure P1 in the first oil chamber 18 is
It is larger than 2 (see FIG. 3). At the same time, the high-pressure oil is supplied to the first oil chamber 18, and the high-pressure oil is discharged and collected from the second oil chamber 19. In this way, the counterweight 15a is moved, and when the position sensor 20 confirms that the counterweight 15a has reached a predetermined position in the circumferential direction of the rotor 1, the movement of the counterweight 15a is stopped. In this case, by stopping the supply and discharge of the high pressure oil by the hydraulic unit 9 and making the hydraulic pressure P1 in the first oil chamber 18 and the hydraulic pressure P2 in the second oil chamber 19 equal, the movement of the counterweight 15a can be prevented. It can be stopped and the counterweight 15a can be held at a predetermined position in the circumferential direction of the rotor 1 by centrifugal force.

Further, the first oil chamber 1 of the counterweight moving pipe 12a
8 and a second oil chamber 19, the hydraulic unit 9 of the hydraulic control device 7 is connected via an oil coupling device _3a 1, 3a _2, oil coupling device _3a 1, 3a ₂ of the oil seal 4 is oil joint Since the sliding contact between the devices 3a ₁ and 3a ₂ and the rotor 1 can be maintained, the balance weight 15a is controlled by controlling the hydraulic pressure from the outside of the rotor 1 while the rotor 1 is rotating.
Can be moved. Similarly to the counterweight 15a, the counterweight 15b can be moved to a predetermined position in the circumferential direction of the rotor 1 while the rotor 1 is rotating and can be held at the predetermined position.

Next, another embodiment of the present invention will be described with reference to FIG. Incidentally, in the second embodiment shown in FIG.
The same components as those in the first embodiment shown in FIGS. 1 to 6 are designated by the same reference numerals, and the description thereof will be omitted. Further, in FIG. 7, the same components as those of the first embodiment are designated by the hydraulic control device 7
It is omitted for convenience except the part.

In FIG. 7, an automatic control signal generator 20 is connected to the control device 8 of the hydraulic control device 7, and a storage device 21 and an arithmetic unit 22 are connected to the automatic control signal generator 20. In addition, the automatic control signal generator 2
To 0, a tachometer (not shown) for measuring the rotation speed of the rotor 1 is connected. A vibration sensor 23 that detects the vibration of the rotor 1 and outputs a vibration signal is connected to the arithmetic unit 22. Further, the storage device 21 stores the influence coefficient vector for all the rotation speeds of the rotor 1 measured in advance.

Next, the operation of this embodiment having such a configuration will be described. During operation of the rotating machine, the vibration sensor 23 detects vibration of the rotor 1 and outputs a vibration signal to the arithmetic unit 23. The vibration signal input to the arithmetic unit 23 is converted into a vibration vector by the arithmetic unit 23,
This vibration vector is input to the automatic control signal generator 20. The automatic control signal generator 20 retrieves the influence coefficient vector corresponding to the rotation speed of the rotor 1 from the storage device 21,
Based on the influence coefficient vector and the vibration vector, a predetermined position in the circumferential direction of the rotor 1 of the balancing weights 15a and 15b required for balancing the rotor 1 is calculated by the above-described influence coefficient method. Then, the automatic control signal generator 2
0 outputs an automatic control signal for moving the balancing weights 15a and 15b to the calculated predetermined position in the circumferential direction of the rotor to the control device 8 of the hydraulic control device 7.

As described above, according to this embodiment, even when the vibration state changes due to the change in the rotation speed of the rotor 1 during the operation of the rotating machine, the rotor 1 is automatically balanced and rotated in response to the change. Vibration of the machine can be reduced.

[0038]

As described above, according to the present invention, the hydraulic pressure in the counterweight moving pipe is controlled by the hydraulic control device to move the counterweight in the weight moving pipe in the rotor circumferential direction,
The counterweight can be moved into position for counterbalancing the rotor of the rotating machine. Therefore, the rotor can be balanced in response to changes in the vibration state without stopping the operation of the rotating machine, and the vibration of the rotating machine can be reduced. Therefore, it is possible to realize vibration reduction of a rotating machine such as a steam / gas turbine used in a thermal power / nuclear power plant, which is likely to change in vibration state, without stopping the operation of the plant. Stable operation can be enabled.

[Brief description of drawings]

FIG. 1 is a diagram showing an embodiment of a balancing device for a rotary machine according to the present invention.

FIG. 2 is a sectional view taken along the line II-II of the balancing machine for the rotary machine shown in FIG.

FIG. 3 is an enlarged cutaway view of the periphery of the counterweight of the rotary machine balancing device shown in FIG.

FIG. 4 is a schematic diagram showing the relationship between the position of the counterweight and the unbalance vector (correction amount) in the balancing machine for the rotary machine shown in FIG.

FIG. 5 is a vibration vector diagram showing an example of an influence coefficient vector and a vibration vector of a rotating machine.

6 is a vector diagram showing a relationship between an unbalance vector (correction amount) for balancing the vibration vector shown in FIG. 5 and a position of a counterweight.

FIG. 7 is a schematic view partially showing another embodiment of the balancing machine for a rotary machine according to the present invention.

FIG. 8 is a side view schematically showing an example of a conventional rotary machine.

FIG. 9 is a diagram showing an example of a change in amplitude with respect to a change in rotation speed of a conventional rotating machine.

FIG. 10 is a vibration vector diagram showing an example of a vibration vector change characteristic with respect to a load change of a conventional rotary machine.

FIG. 11 is a vibration vector diagram showing another example of a vibration vector change characteristic with respect to a load change of a conventional rotating machine.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 rotor 3 balancing device 9 hydraulic control device 12a, 12b balancing weight moving pipe 15a, 15b balancing weight 20 control signal generator 21 memory device 22 arithmetic unit 23 vibration sensor

Claims (2)

[Claims] 1. A ring-shaped counterweight moving pipe provided on the outer circumference of a rotor of a rotating machine, and a ring-shaped counterweight moving pipe housed in the counterweight moving pipe and movable in the circumferential direction of the rotor by hydraulic pressure in the counterweight moving pipe. And a hydraulic control device for controlling the hydraulic pressure in the counterweight moving pipe to move the counterweight to an arbitrary position in the rotor circumferential direction. Balancing device for rotating machinery. 2. A storage device for storing an influence coefficient vector of vibration at all rotation speeds of the rotor, a vibration sensor for detecting vibration of the rotor and outputting a vibration signal, and detecting a rotation speed of the rotor. An arithmetic unit for converting the vibration signal from the tachometer and the vibration sensor into a vibration vector, and based on the vibration vector from the arithmetic unit and the influence coefficient vector of the storage device corresponding to the rotation speed from the tachometer, A control signal that calculates a predetermined position in the rotor circumferential direction of the counterweight necessary to balance the rotor and outputs a signal for moving the counterweight to the predetermined position to the hydraulic control device. The balancing device for a rotary machine according to claim 1, further comprising: a generator.