JPH09133127A - Pad-type journal bearing - Google Patents

Abstract

PROBLEM TO BE SOLVED: To avoid harmonic resonance based on the rocking vibration of a pad and to make the vibration of a pad-type journal bearing sufficiently stable. SOLUTION: This pad-type journal bearing has a cylindrical bearing inner ring 21 and at least one circular pad 22 having a spherical back face 22 b. The pad 22 is rockingly disposed in a state in which the back face 22b is in contact with the inner circumferential surface 21a of the bearing inner ring 21 and the rocking vibration of the pad 22 makes the oil film pressure of lubricating oil change to support the journal of a rotating shaft in automatic aligning state. The back face 22b is formed in a radius of curvature Ria as a means for avoiding harmonic resonance such that the frequency of rocking vibration of the pad 22 is N and 1/N times rotation frequency (where N is a natural number of 1 or more) plus a prescribed margin, or off a harmonic frequency.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a pad type journal bearing for supporting a rotary shaft of a large rotary machine such as a steam turbine / generator, and is particularly applied to a high speed rotary machine that requires high stability. The present invention relates to a pad type journal bearing.

[0002]

2. Description of the Related Art Conventionally, as shown in FIG. 21, a pad type journal bearing of this type has a cylindrical bearing inner ring 1 having a curved inner surface 1a, and a sheet surface (in a plane orthogonal to a rotation axis).
And a plurality of pads 2 provided in contact with the inner peripheral surface (inner surface) 1a of the bearing inner ring 1 so as to be swingable in the direction perpendicular to the plane of the drawing.
2, and the journal portion (journal) 3 of the rotating shaft is inserted and supported inside the bearing inner ring 1 and the pad 2. The bearing inner ring 1 is fixedly supported from the outside by a cylindrical bearing outer ring 4 as well, and the bearing outer ring 4 is fixedly mounted on a bearing base (casing or foundation) 5 by, for example, bearing mounting bolts 6 or the like. There is.

The plurality of pads 2 are arranged concentrically around the central axis of the journal 3. Each pad 2 is formed as an arc-shaped piece having a predetermined width. That is, the surface (inner surface) 2a of the pad 2 facing the journal side is a curved surface that is formed by a predetermined curvature radius, and the outer surface (back surface) 2b that faces the inner surface is also formed by the same curvature radius as the inner surface. Is a curved surface. Each pad 2 is an inner surface 1 of the bearing inner ring 1.
It is provided with its back surface 2b in contact with a.

A fixing hole 7 is formed in the back surface 2a of the pad 2 by processing and a supporting pin 8 is inserted through the fixing hole 7 through the bearing inner ring 1 and the bearing outer ring 4. That is, there is a gap between the support pin 8 and the fixing hole 7, and the pad 2 is swingably supported by the support pin 8 while maintaining a margin for the gap. Further, the inner surface 2a of the pad 2 is formed by bonding a bearing alloy such as white metal to improve lubrication characteristics.

On the other hand, lubricating oil is supplied into the bearing inner ring 1 through an oil supply line 9 provided in the bearing inner ring 1. This lubricating oil flows through the gap 10 between the journal 3 and the pad 2, and is then discharged from the bearing through the oil drain port 11. The lubricating oil flowing in the gap 10 of the journal bearing has an oil film pressure F _p distributed as shown in FIG. 22 due to the wedge effect generated in the narrow gap between the inner surface 2a of the pad 2 and the outer peripheral surface of the journal 3 when the journal 3 rotates. The generated journal 3 of the rotating shaft is supported via this oil film pressure F _p . When the oil film pressure distribution F _p is integrated, it is balanced with the bearing load P, and the pad 2 has a contact point T 1 between the back surface 2 b and the bearing inner ring inner surface 1 a of the oil film pressure distribution F _p that changes according to the rotation of the journal 3. Move freely to be located vertically below the center point. Therefore, the pad 2 is adapted to receive the bearing load P in a self-centering manner as a whole.

A characteristic of the pad type journal bearing is the self-centering function. That is, the pad-type journal bearing has a self-aligning property that can follow the vibration of the rotating shaft, and thus has a very stable mechanism as compared with other journal bearings.

The above-mentioned automatic centering function will be specifically described below.
As shown in FIGS. 22 and 23, the center O of the journal 3
_{Even when j} moves to O _{j ′} due to vibration, the oil film pressure distribution F _p changes with the movement of the center of the journal 3, so that the pad 3 is placed vertically below the center of the pressure distribution and the pad back surface 2 b and the bearing inner ring. The rocking motion is performed so that the contact point with the inner surface 1a is located. As a result, as shown in FIG. 23, the contact point between the pad back surface 2b and the bearing inner ring inner surface 1a is from T1 to T1.
Go to 2. As described above, since the pad freely moves according to the vibration of the journal 3, it is possible to prevent the occurrence of the destabilizing force that causes the whip or the like in the bearing, and to supply the journal bearing with excellent stability. Become.

[0008]

However, due to the recent increase in capacity of steam turbines and generators, the size of the rotating body or the bearing that supports it has also been increasing. Due to this increase in size, the rotating shaft / bearing system is becoming susceptible to vibration, and asynchronous or unstable vibration that does not synchronize with rotation also occurs in the rotating shaft system that uses a highly stable pad-type journal bearing. It was way.

FIG. 24 is a diagram showing vibrations that have occurred in a large steam turbine in the past. The horizontal axis represents time, and the vertical axis represents load and vibration amplitude of the steam turbine. It is a thing. From this figure, it can be seen that as the load is increased, the vibration amplitude value increases, and it is also recorded with a width on the vibration chart, which is a characteristic of asynchronous vibration. The rotation-synchronized vibration that occurs due to normal rotor imbalance or the like is stable, and appears on the vibration chart as a single line with no width, such as the vibration value A _{m1 in} FIG. However, since the asynchronous vibration itself has non-stationarity, the vibration amplitude generally changes with time in a short cycle, and the vibration amplitude value generally appears on the vibration chart as a line A _m2 having a width. is there.

In the past, the cause of these asynchronous vibrations was not clear, but due to recent advances in vibration analysis technology, such asynchronous vibrations generated in pad-type journal bearings are non-linear vibrations due to oscillating vibrations of the pad. It has been found that the vibration of the journal acts as an exciting force on the pad due to the above, and causes harmonic resonance such as subharmonic or multiple harmonic. The mechanism and characteristics of the vibration will be described below.

FIG. 23 described above is also a diagram for explaining the mechanism of generation of the asynchronous vibration. As described above, the movement of the pad follows the vibration of the journal, and the contact point T1 on the back surface of the pad is displaced to the new contact point T2. The movement of the contact point is represented by a displacement angle θ from the center point of the inner surface 1a of the bearing inner ring as shown in FIG. 24, and the equation of motion is represented by the following equation (1) using various elements of the pad.

[0012]

(Equation 1) It is.

FIG. 25 shows an example of the response analysis of equation (1).
On the horizontal axis, the vibration frequency f of the vibration force F with respect to the oscillation natural frequency f _n of the pad (mainly corresponds to the rotation speed f _o of the shaft on the rotation axis).
Frequency ratio f / f _n , and the vertical axis represents the movement angle of the contact point between the pad back surface and the bearing inner ring in FIG. 24. The X coordinate axis (vertical Y coordinate axis in the swing plane of the pad 2 is perpendicular to the X coordinate axis). The horizontal axis represents the response characteristics when the vibration displacement is converted.

Equation (1) shows the equation of motion for the oscillating vibration of the pad. This equation is a typical non-linear equation of motion of a variable elastic system, and its response also has characteristics peculiar to the non-linear system. It should be shown. As can be seen from FIG. 25,
First, resonance occurs at the frequency ratio f / f _n = 1/2.
At this time, since the vibration frequency of the pad matches the oscillation natural frequency, it is twice the vibration frequency, which is called double resonance. The phenomenon of resonating at a frequency N times as high as the excitation frequency (N is a natural number of 1 or more) is called Nth harmonic resonance. From the figure, the frequency ratio f / f
It can be seen that resonance occurs at _n = 1. At this time,
Since the vibration frequency and the excitation frequency match, this is called harmonic resonance. Further, it can be seen that a small resonance occurs at the vibration ratio f / f _n = 2/3. This is a resonance peculiar to nonlinearity, and is a phenomenon in which the system resonates when the vibration frequency ratio matches n / m (both n and m are natural numbers of 1 or more), and is called high subharmonic resonance.

Further, it can be seen that resonance occurs at the frequency ratio f / f _n = 2. At this time, the vibration frequency is equal to 1/2 of the vibration frequency because it matches the oscillation natural frequency, which is called a 1 / 2-order subharmonic resonance.
The phenomenon of resonating at a frequency that is 1 / N times the vibration frequency (N is a natural number of 1 or more) is called Nth-order subharmonic resonance. Hereinafter, these resonance phenomena will be collectively referred to as harmonic resonance.

As described above, in recent years, the diameter of the journal bearing has tended to increase due to the increase in single machine output of the steam turbine and the improvement in performance, and the pad size naturally increases as the journal bearing diameter increases. As the pad becomes larger, the rocking natural frequency of the pad tends to decrease.

For example, the rocking natural frequency f _{n of the} pad is 1
If it is at 00 Hz, 1/2 of that, 50 Hz, just hits the operating speed, and asynchronous resonance occurs as shown in FIG. 25. The vibration example of FIG. 24 is generated by this 1/2 subharmonic resonance. This vibration exhibits non-stationarity peculiar to a non-linear phenomenon and has a vibration characteristic in which the amplitude changes with time.

As described above, the oscillating vibration of the pad, which is the cause of the asynchronous signal, causes harmonic resonance, and when this harmonic resonance occurs, the oscillating vibration of the pad is transmitted through the lubricating oil film of the bearing to the rotating shaft. , And the rotating shaft vibrated and had an adverse effect. The adverse effects of the oscillation vibration and the harmonic resonance of the pad are not based on the unstable vibration that causes the rotation shaft system to be destroyed like the oil whip of the bearing that has been conventionally known, but the pad with good stability is provided. In providing the type journal bearing, it became necessary to avoid harmonic resonance (subharmonic vibration / multiple harmonic vibration) related to the oscillation vibration of the pad.

The present invention has been made in view of the above circumstances, and an object of the present invention is to avoid harmonic resonance due to rocking vibration of a pad and to sufficiently stabilize the pad type journal bearing in terms of vibration.

[0020]

As described above, many resonance phenomena occur in a non-linear system such as rocking vibration of a pad, which is N times or 1 / N times the rocking natural frequency of the pad. It is known that resonance (harmonic resonance) is exhibited at the frequency of the excitation force such as. Then, as shown in FIG. 25, this harmonic resonance occurs according to the ratio of the oscillation natural frequency of the pad to the rotational speed of the journal.

Therefore, in order to avoid the occurrence of many harmonic resonance phenomena in a nonlinear system such as rocking vibration of the pad,
The oscillation frequency of the pad is N times the rotational speed used, 1
It is understood that it is good if it is not near the frequency such as / N times.

That is, the present inventor prevents the occurrence of the above harmonic resonance phenomenon by controlling the oscillation natural frequency of the pad to a value other than near the frequency such as N times or 1 / N times the rotational frequency used. Focused on what is possible.

Here, the rocking natural frequency f _n of the pad is
When the right side of the equation (1) is set to "0" and the equation of motion regarding the free vibration is solved, it is expressed as the equation (2).

[0024]

(Equation 2)

As can be seen from this equation, the oscillation natural frequency of the pad changes depending on various factors such as the geometrical shape of the pad and its mass. That is, (1) change in the mass of the pad, (2) C _r (change in difference in curvature between the inner curved surface of the bearing and the back surface of the pad; hereinafter referred to as pad gap), (3) change in curvature of the back surface of the pad. , (4) changes in rotational rigidity k _θ , and (5) changes in bearing load can change the oscillation frequency f _n of the pad.

The first invention of the present application created based on the above process is as described in claims 1 and 2.
A cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface, and the pad is swingably disposed with the back surface in contact with the inner peripheral surface of the bearing portion, In a pad type journal bearing in which the oil film pressure of the lubricating oil is changed by the oscillation vibration of the pad to support the journal portion of the rotating shaft in a self-aligning manner, the frequency of the oscillation vibration of the pad is set to the rotation of the journal portion. A detuning means for detuning from a value of N times and 1 / N times the frequency (N is a natural number of 1 or more) with a predetermined margin is provided.

In particular, in the first aspect of the invention, as the detuning means, the back surface of the pad has a predetermined oscillation vibration frequency of N times and 1 / N times the rotational frequency. It is formed with a radius of curvature that provides a frequency that is detuned with a margin, and the vibration frequency of the rocking vibration of the pad is N times the rotational frequency at the position where the pad contacts the inner peripheral surface of the bearing portion. And a pad contact surface formed with a radius of curvature that is a frequency that is detuned with a predetermined margin from the value of 1 / N times.

Further, in the first aspect of the present invention, in particular, as the detuning means, the mass of the pad is set so that the oscillation frequency of the pad is N times or 1 / N times the rotational frequency. It is set so that the frequency is detuned with a predetermined margin. For this mass adjustment, for example, the pad is provided with a required number of holes having a required shape and a required length along the axial direction of the journal.

Further, in the first aspect of the invention, as the detuning means, the bearing load applied to the pad is determined such that the oscillation frequency of the pad is N times or 1 / N times the rotational frequency. It is set so that the frequency is detuned with a predetermined margin. In order to set the bearing load to the detuned frequency, the fixed support position of the bearing portion is changed to adjust the bearing alignment of the bearing portion to a required value.

Further, in the first invention, as the detuning means, a rotational rigidity imparting means for imparting a predetermined rotational rigidity to the pad is inserted between the pad rear surface and the inner peripheral surface of the bearing portion. Then, the vibration frequency of the rocking vibration of the pad is detuned from the values of N times and 1 / N times the rotation frequency with a predetermined margin according to the rotation rigidity given from the rotation rigidity imparting means. Set to the frequency. For example, the rotational rigidity imparting means is an elastic member such as a coil spring having a predetermined elastic force.

According to the first aspect of the invention, as the detuning means, for example, on the back surface of the pad, the oscillation frequency of the pad is set to a predetermined value from N times and 1 / N times the rotational frequency. By forming with a radius of curvature that provides a frequency that is detuned with a margin, the vibration frequency of the rocking vibration of the pad is set to a value that is N times or 1 / N times the rotational frequency of the journal portion, for example. Since detuning is performed with a predetermined margin, it is possible to prevent the occurrence of a harmonic resonance phenomenon. Note that detuning the oscillation frequency of the pad with respect to the values of N times and 1 / N times the number of rotations used means that the oscillation frequency of the pad is N times and 1 / N times the number of rotations used. It means setting to a value that is off (remote) from.

Further, the present inventor has conducted extensive research on means for suppressing the harmonic resonance phenomenon even when the harmonic resonance phenomenon occurs. And when the harmonic resonance phenomenon occurs,
It was found that when the vibration system is damped, the vibration amplitude is suppressed.

FIG. 26 shows a change in the characteristic when the response calculation is performed by adding damping to the vibration characteristic described in FIG. When the damping ratio ζ of the system with respect to the rocking vibration of the pad is selected as a parameter and the value of ζ is increased from 0 → 0.05 → 0.1, the vibration amplitude just decreases as shown in FIG. Or, "1/2 subharmonic resonance", "2
It can be seen that "/ 3 subharmonic resonance" has disappeared.

As described above, by providing damping to the vibration system, even if a harmonic resonance phenomenon occurs due to insufficient detuning of the oscillation natural frequency of the pad, the vibration amplitude of the harmonic resonance can be reduced. it can.

According to a second invention of the present application created based on the above process, as described in claim 24, a cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface are provided. The pad is swingably arranged with its back surface in contact with the inner peripheral surface of the bearing portion, and the lubricating oil film pressure is changed by the swing vibration of the pad to automatically rotate the journal portion of the rotating shaft. In a pad type journal bearing that is supported in an aligned manner, when a harmonic resonance phenomenon occurs in the pad,
A damping mechanism for damping the vibration amplitude of the pad due to the harmonic resonance is provided.

In particular, as a second aspect of the invention, the damping mechanism is provided at the back end of the pad so that the radius of curvature of the pad back end is substantially the same as the radius of curvature of the inner peripheral surface of the bearing portion. It has a squeeze section provided.

According to the second invention, even when a harmonic resonance phenomenon occurs in the pad, the radius of curvature of the back end of the pad and the radius of curvature of the inner peripheral surface of the bearing part are reduced by, for example, a squeeze portion as a damping mechanism. Since they are almost the same, the squeeze effect is efficiently generated in the lubricating oil that has flowed into the gap between the pad back end and the inner peripheral surface of the bearing, and this squeeze effect causes vibration of the pad due to the harmonic resonance. The amplitude can be attenuated.

Further, the inventor of the present invention suppresses the vibration of a vibration system such as the oscillating vibration of a pad by actively applying an exciting force for suppressing the vibration to the vibrating object. Focusing on the fact that it is possible, we created the third invention.

That is, a third aspect of the present invention comprises, as described in claim 28, a cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface, and the pad having the back surface thereof. A pad-type journal that is swingably arranged in contact with the inner peripheral surface of the bearing portion and that changes the lubricating oil film pressure by the swing vibration of the pad to support the journal portion of the rotating shaft in a self-aligning manner. In the bearing, when a harmonic resonance phenomenon occurs in the pad, active control means for actively controlling the vibration of the pad due to the harmonic resonance to a predetermined value that does not affect the rotational vibration of the journal portion is provided. There is.

In particular, the active control means is provided at least in the vicinity of one end of the back surface of the pad, and detects the vibration displacement of the pad, and the active control target is based on the displacement data detected by the detection means. Generating means for generating vibration information, an oil supply line communicating with at least a gap between the other end of the back surface of the pad and the inner peripheral surface of the bearing portion, a supply means for supplying lubricating oil to the oil supply line, and the oil supply line And a valve opening / closing timing control means for controlling the opening / closing timing of the valve according to the vibration information.

According to the third aspect of the invention, as the active control means, for example, the detecting means detects the vibration displacement of the pad, and the generating means generates the vibration information for the active control based on the detected data. On the other hand, as an active control means, an oil supply line communicating with the gap between the other end of the back surface of the pad and the inner peripheral surface of the bearing, a supply means for supplying lubricating oil to this oil supply line, and an oil supply line inserted in the middle of the oil supply line A valve for controlling the opening and closing of the pad is provided, the valve opening and closing timing control means controls the opening and closing of the valve at a predetermined timing according to the vibration information, and the lubricating oil is supplied to the other end of the pad back surface through the supply means and the oil supply line. It is supplied to the gap with the inner peripheral surface of the bearing. That is, since the lubricating oil is supplied to the gap according to the vibration displacement of the pad, the vibration of the pad due to the harmonic resonance is actively controlled to a predetermined value that does not affect the rotational vibration of the journal portion. Therefore, the vibration due to the harmonic resonance can be suppressed.

[0042]

Embodiments of the present invention will be described below with reference to the accompanying drawings. In all of the embodiments described below, only the essential part of the present invention is described, and the overall structure of the pad type journal bearing is shown in FIG.
Since it is substantially the same as the configuration shown in FIG. Therefore, the description of each component of the entire pad type journal bearing, the journal supporting action by the oil film of the lubricating oil, and the self-centering action due to the change of the oil film pressure distribution are substantially the same as those described in the conventional example. Therefore, it is omitted here.

Further, as described above, according to the first invention of the present application, various parameters for determining the oscillation natural frequency of the pad expressed by the above equation (2), that is, the radius R of the back surface of the pad.
By adjusting _i , pad clearance C _r , pad mass m, bearing load P, pad rotational rigidity k _θ , the oscillation natural frequency of the pad is N times and 1 / N times (N is This is to prevent harmonic resonance by performing detuning with a margin with respect to a value of 1 or more). Note that detuning the oscillation frequency of the pad with respect to the values of N times and 1 / N times the number of rotations used means that the oscillation frequency of the pad is N times and 1 / N times the number of rotations used. Setting to a value outside the range.

The second aspect of the present invention is to reduce the vibration amplitude of the harmonic resonance by damping the vibration system. Further, a third aspect of the present invention is to suppress the vibration by actively applying a vibrating force for damping the vibration to the swing vibration system of the pad.

Therefore, the respective embodiments of the first invention, the second invention, and the third invention will be described below.

First, the means for adjusting various parameters in the first invention will be described in each embodiment.

(First Embodiment) In the present embodiment, by adjusting the difference between the radius of curvature of the back surface of the pad and the radius of curvature of the inner surface of the bearing inner ring (hereinafter, this difference is referred to as the pad gap), the vibration of the pad is adjusted. Dynamic natural frequency is N times the rotation speed and 1
A means for performing detuning with a margin for frequencies such as / N times will be described.

FIG. 1 is a front view showing a pad which is a main part of the pad type journal bearing of the first embodiment. In FIG. 1, the inner surface 21a of the cylindrical bearing inner ring 21 is molded with a radius of curvature R _O (referred to as the center of curvature O _{b of the} inner surface of the bearing inner ring). The pad 22 is an arc-shaped piece having a predetermined width,
The surface (inner surface) 22a of the pad 22 that faces the journal side is a curved surface that is molded with a predetermined radius of curvature R _b . Further, in the present embodiment, the inner surface 22a facing the outer surface (back surface) 22b is (a pad back curvature center O _p) a radius of curvature R _i is molding such that the required value R _ia different from the inner surface 22a ing. The pad 22 has a back surface 22
b is arranged in contact with the inner surface 21a of the inner surface 21a of the bearing inner ring.

The predetermined value R _{ia of the} radius of curvature of the back surface 22b of the pad 22 is a pad gap C _rA which changes based on the R _{ia .}
(= R _o −R _ia ) is set to have a predetermined length, and the predetermined length of the pad gap C _rA is the pad 22.
Oscillation frequency of N is the number of rotations used (the number of rotations of the shaft) fo
Fold and 1 / N times {eg half (1/2 · f _o), 1
Double (f _o ), double (2 · f _o ), etc.} is detuned with a margin.

Further, FIG. 2 shows that when the radius of curvature R _i of the back surface 23b of the pad 23 (the center of curvature of the pad back surface is O _p ) is uniformly set as in the conventional case, the radius of curvature R _o (bearing A pad contact with a radius of curvature _Roa is newly centered on O _ba in the range where the back surface 23 b of the pad 23 of the inner surface 24 a of the bearing inner ring 24, which is molded at the center of curvature O _b of the inner surface of the inner ring, contacts. A surface 24b is provided.

The radius of curvature R _oa pad abutment surface 24b, the pad clearance _{C rB (= R oa -R i} ) has been determined to be a predetermined length, the predetermined length of the pad clearance C _rB That is, the oscillation natural frequency of the pad 23 is N times and 1 / N times the used rotation speed (rotation speed of the shaft) fo (for example, 1/2 times (1 /
2 · f _o ), 1 × (f _o ), 2 × (2 · f _o ), etc.} is detuned with a margin.

Here, in FIG. 3, the pad gap C _r, that is, the difference between the inner surface radius of curvature R _{o of the} bearing inner ring and the curvature R _i of the pad back surface is plotted on the abscissa, and the ordinate of the pad is calculated from equation (2). It shows a change in frequency characteristic when the dynamic natural frequency is taken. As shown in FIG. 3, the relationship between the gap C _r and the oscillation natural frequency f _n is a parabola that descends to the right. When the gap C _r is reduced, the induced natural frequency f of the pad is reduced.
It can be seen that _n gradually decreases.

In order to prevent the harmonic resonance of the fractional / multiple harmonic vibration which is the nonlinear resonance as shown in FIG. 25, the oscillation natural frequency of the pad is N times and 1 / the rotational speed.
It is important not to be N times (particularly 1/2 times), and
Since the purpose of this embodiment is to avoid the resonance response, a certain margin must be set for the frequency to be avoided. This margin, which is determined by the damping of the oscillating vibration, is about ± 10%.

For example, at the number of revolutions f _o used, its frequency f _o
To as twice a multiple harmonic resonance is not generated, it will be rocking natural frequencies of the pad 2 times the value (2 · f _o) of the operational rotational speed f _o may If not, further If a margin of ± M% (for example, ± 10%) is taken in consideration of its responsiveness, the oscillation natural frequency becomes a value that satisfies the following conditional expression (3) in expression (2). Thus, the pad gap C _r may be determined.

Now, in equation (2), if parameters other than the pad gap C _r are constants,

(Equation 3) (A and B are constants) and the pad gap C _{r is the} only equation.

(Equation 4) It suffices to define C _r that satisfies Then, the pad back surface 20b is formed so as to satisfy the determined clearance C _r.
The radius of curvature R _ia (FIG. 1) or of the pad contact surface 21b of curvature radius R _oa (Figure 2) by molding, it is possible to avoid the double multiple harmonic resonant frequency f _o.

[0056] Incidentally, the hatched portion in FIG. 3, when the rocking natural frequencies in the range of _{"f o, 1/2 · f} o, 2 · f o " and its marginal portion, the rocking unique The range of the gap C _r corresponding to the range of frequencies is shown. That is, the relationship of the equation (3) indicates a frequency range other than the frequency range (shaded portion) of the oscillation natural frequency “2 · f _o ” including the margin portion in FIG.

On the other hand, in order to avoid harmonic resonance other than the multiple harmonic resonance, a conditional expression similar to the expression (3) is added for each frequency. Other procedures for avoiding the multiple harmonic resonance will be described with reference to FIG. That is, as shown in FIG. 3, for example, in order to avoid second harmonic resonance, the oscillation natural frequency should not be in the vicinity of 2 · f _o . Therefore, if the gap C _r (the radius of curvature R _ia of the pad back surface 22b or the radius of curvature R _oa of the pad contact surface 24b) is set to a length other than the range of the gap (gap C _{r1 to} C _r2 ) shown by the diagonal lines. Good. Further, in order to avoid 1,1 / 2 times the harmonic resonance, rocking natural frequencies must not be in the vicinity of _{f o, 1/2 · f} o. Therefore, the gap C _r has a length other than the shaded gap range (gap C _{r3 to} C _r4 , C _{r5 to} C _r6 ).
Should be set. In this _{C r1 ~C r2, C r3 ~C} r4 ,, and C _r5 set the range in the gap C _r of -C _r6, rocking natural frequencies of the pad, the used rotation speed f _o 1/2 Double (1 /
2 · f _o), 1 times (f _o), will be detuned with a margin with respect to the frequency of such as a two-fold (2 · f _o).

Therefore, as in the case of multiple harmonic resonance, FIG.
In the above configuration, C _{r1 to} C _r2 , C _{r3 to} C _r4 , and C _{r5 to} C _r6.
The radius of curvature R _ia of the pad back surface 22b may be molded so that the clearance C _r is outside the range of C _r1 , and in the configuration of FIG. 2, C _{r1 to} C _r2 , C _{r3 to} C _r4 , and C _{r. r5} so that outside of the gap C _r of -C _r6, the radius of curvature R _oa pad contact surface 24b may be molding.

[0059] As described above, according to this embodiment, by adjusting the curvature radius of the curvature radius or pad contacting surface of the pad back, the swing natural frequency of the pad operational rotational speed f _o 1 / 2 times (1/2 · f _o), 1 times (f _o),
Since it is possible to detune with double frequency (2 · f _o ) with a margin, harmonic resonance due to non-linear vibration of the pad is avoided, and pad type journal bearing with improved stability is provided. can do. In the above embodiment, the rocking natural frequencies of the operating rotation speed f _o of the pad 1 /
2 times (1/2 · f _o), 1 times (f _o), 2-fold (2 ·
Although an example in which detuning is performed with a margin with respect to f _o ) has been shown, the present invention is not limited to this.
By using the same idea as in FIG. 3, it is possible to detune the oscillation natural frequency of the pad with a margin with respect to N times or 1 / N times the number of revolutions. can get.

Further, generally, each pad of the pad-type journal bearing is not uniformly loaded, and the rocking natural frequency of the pad changes depending on the load P applied to the pad as shown in the equation (2). Therefore, the rocking natural frequencies of the pads arranged concentrically forming the pad journal bearing are different from each other. Therefore, in this case, half of the swing characteristic frequency corresponding to the load of each pad using rotational speed _{f o (1/2 · f o)} , 1 -fold (f _o), 2-fold ( The frequency may be detuned with a margin for frequencies such as 2 · f _o ). That is, the pad gap C _{r for} each pad is equal to the frequency f _o , 1 / 2f
C _{r1 to} C _r2 and C _r3 corresponding to _o , f _o and 2 · f _o
The radius of curvature R _ia of the back surface of each pad may be molded so that it is out of the range of C _r4 and C _{r5 to} C _r6 .

FIG. 4 is a diagram showing the pads 25a to 25d molded as described above. As can be seen from the equation (2), when the rocking natural frequency of the pad is kept constant, if the load applied to the pad is large, the pad gap C _r must be increased accordingly. Usually, as shown in FIG. 4, a load of P1 (vertically downward) is applied to the pad 25a and a load of P2 (one side direction) is applied to the pad 25b.
The load of P3 (vertically upward) on the pad 25d and P2 on the pad 25d.
A load (in the direction of the other side) is applied, and its size is P
1>P2> P3. Therefore, in order to detune the rocking natural frequency of each of the pads 25a to 25d as shown in FIG. 3, the respective pads must be detuned. Figure 4 is a half the swing natural frequency of the operating rotation speed f _o of each pad (1/2 · f _o), 1-fold (f _o), a margin from the double (2 · f _o) Each pad 25a is designed to be held and detuned and set to the frequency f _p.
˜25d has different radii of curvature on its back surface, such as Ri1, Ri2, Ri3, and Ri2, and its size is
It is formed so that Ri1 <Ri2 <Ri3.

The pad type journal bearing shown in FIG.
Since the oscillation natural frequency of each pad can be accurately detuned with a margin for frequencies of N times and 1 / N times the number of rotations used, tuning based on the nonlinear vibration of the pad can be performed. It is possible to provide a pad-type journal bearing that is more effective in avoiding wave resonance and has improved stability as compared with the pad-type journal bearing having the configuration shown in FIGS. 1 and 2.

(Second Embodiment) In the present embodiment, by adjusting the mass m of the pad, the oscillation natural frequency of the pad has a margin for frequencies such as N times and 1 / N times the number of rotations used. The means for holding and detuning will be described.

FIG. 5 shows changes in frequency characteristics when the mass m of the pad 26 is plotted on the abscissa and the oscillation natural frequency of the pad 26 obtained on the ordinate is obtained from the equation (2). As shown in FIG. 5, when the mass m of the pad 26 is changed, the rocking natural frequency f _n is also changed according to the change. That is, as shown in FIG.
It is understood that the relationship between the mass m and the oscillation natural frequency f _n is a parabola that descends to the right, and the induced natural frequency f _n of the pad 26 gradually decreases as the mass m decreases. .

For example, at the number of revolutions f _o used, its frequency f _o
In order to prevent generation of a harmonic resonance twice as high as that of the pad, it is sufficient that the oscillation natural frequency of the pad does not exist at a value (2 · f _o ) that is twice the rotational frequency f _o used. Therefore, as shown in FIG. 5, the mass m of the pad is determined within a range other than the range (m1 to m2) of the shaded part corresponding to the frequency range of the oscillation natural frequency “2 · f _o ”, including the margin part. Just do it.

As described above, the above-mentioned pad gap C
Similarly to setting _r (the radius of curvature R _ia of the pad back surface 22b or the radius of curvature R _oa of the pad abutment surface 24b), by adjusting the mass m of the pad 26, the oscillation natural frequency is changed to the rotation frequency. 1/2 of _{f o (1/2 · f o)} , 1 times (f _o), can be detuned with a margin with respect to the frequency of such as a two-fold (2 · f _o).

In order to adjust the mass m of the pad 26, changing the material of the pad is the most obvious method, and the mass m can be reduced or increased. However, by using machining, adjustments can be easily made even with existing pads made of normal materials.

FIG. 6 shows a pad mass m obtained by the above machining.
FIG. 6 is a view of the pad 26 as viewed from the back side to show an example of the reducing means of FIG. According to FIG. 6, a plurality of lateral holes 27 ... 27 are respectively formed at both ends of the pad 26 in the longitudinal direction by machining from one side surface of the pad 26 in the longitudinal direction toward the other side surface. The mass m of the pad 26 can be adjusted to an appropriate value by appropriately adjusting the hole diameter, the hole length, the number of holes, etc. of the lateral holes 27. As for the oil film pressure distribution generated in the pad 26 by the rotation of the journal 3, the maximum pressure is generated in the central portion of the pad, as shown in FIG. Further, since the center of the pressure distribution is also in the central part of the pad, the contact point between the back surface 26b of the pad 26 and the inner surface of the bearing inner ring is located near the center of the pad 26. Therefore, the central portion of the pad must have a certain thickness and rigidity so as to withstand a high load. However, since the pressure distribution is small at both ends of the pad 26, it can be seen that both ends need not be so rigid. As a result, when the pad is processed to reduce and adjust its mass, it is appropriate to drill holes near both ends. Therefore, as shown in FIG. A lateral hole 27 is machined and installed near the portion to reduce the mass m.

FIG. 7 shows a pad 2 in order to show another example of the means for reducing the pad mass m by the above machining.
6A is a view of 6A as viewed from the back side. FIG. According to FIG. 7, a plurality of vertical holes 27A ... 27A are provided by machining from the back surface side to the inner surface side (along the radial direction of the journal 3) at both ends of the pad 26A in the longitudinal direction. The vertical holes 27A ... 27A do not penetrate to the inner surface of the pad 26A, and are machined so as to leave a sufficient thickness for bearing hydraulic pressure received by the portion where the vertical holes 27A are provided. As in the case of FIG. 6, this vertical hole 27
The mass m of the pad 26A can be adjusted to an appropriate value by appropriately adjusting the hole diameter, the hole length, the number, etc. of A ... 27A.

Further, FIG. 8 shows a pad 26 in order to show another example of the means for reducing the pad mass m by machining.
It is a figure when B is seen from the back side. According to FIG. 8, a plurality of grooves are provided on the back surface 26b of the pad 26B. As shown by 28 ... 28 in FIG. 8, the grooves may be formed in an arc shape from both ends of the one side surface in the lateral direction to both ends of the other lateral surface in the lateral direction, or 29. As indicated by 29, they may be formed over the required length from both ends of both side surfaces in the lateral direction.

By appropriately adjusting the depth of the grooves 28 ... 28 (grooves 29 ... 29) and the number of the grooves, the pads 26B can be formed.
It is possible to adjust the mass m of the to a suitable value. As described above, the grooves 28 and 29 are not provided in the vicinity of the contact point T1 on the back surface of the pad with the inner surface of the bearing inner ring, but are processed and molded with sufficient rigidity to withstand the pressure distribution generated in the pad 25B. .

Furthermore, FIG. 9 shows a pad 2 for showing another example of means for reducing the pad mass m by machining.
It is a front view of 6C. According to FIG. 9, a part (reduced portion) 30 of both end portions on the back surface side of the pad 26C is cut by work molding to reduce the pad mass m. That is, the mass m of the pad 26C can be adjusted to an appropriate value by appropriately adjusting the volume (cutting amount) of the weight reducing portion 30.

Further, FIG. 10 is a front view of a pad 26D for showing another example of the means for reducing the pad mass m by machining, similar to FIGS. 6 to 9. When the pad 26D is manufactured, Inner surface 26a and back surface 2 of the pad 26D
The angle of the central angle θ defining 6b is changed from the normal angle θ1 to θ2.
Change to (θ1> θ2), perform molding, and pad mass m
Is to reduce. That is, by appropriately adjusting the angle θ of the pad formation angle (center angle), the pad 26
It is possible to adjust the mass m of D to an appropriate value.

As described above, according to this embodiment,
By adjusting the machining shown the mass of the pad changes and FIGS. 6 to 10 of the material, half of the operational rotational speed f _o rocking natural frequencies of the pad (1/2 · f _o) , 1 times (f _o ), 2 times (2 · f _o ), etc. can be detuned with a margin, so that harmonic resonance due to non-linear vibration of the pad is avoided and stability is improved. It is possible to provide an improved pad type journal bearing.

As described in the description of FIG. 4,
It is possible to provide a more stable journal bearing by adjusting the mass of each pad individually according to the load applied to each pad or by adjusting the material by machining shown in FIGS. 6 to 10.

(Third Embodiment) In the present embodiment, by adjusting the bearing load P, the oscillation natural frequency of the pad has a margin for frequencies such as N times and 1 / N times the number of rotations used. A means for detuning will be described.

FIG. 11 shows changes in the frequency characteristics when the bearing load P is plotted on the horizontal axis and the rocking natural frequency of the pad, which is obtained from equation (2), is plotted on the vertical axis. As shown in FIG. 11, it can be seen that when the bearing load P is changed, the oscillation natural frequency f _n is also changed according to the change. That is, as shown in FIG. 11, the relationship between the bearing load P and the oscillation natural frequency f _n is an upward-sloping parabola, and as the bearing load P is reduced, the oscillation natural frequency f _{n of the} pad is reduced.
It can be seen that is gradually rising.

For example, at the number of revolutions f _o used, its frequency f _o
In order to prevent generation of a harmonic resonance twice as high as that of the pad, it is sufficient that the oscillation natural frequency of the pad does not exist at a value (2 · f _o ) that is twice the rotational frequency f _o used. Therefore, as shown in FIG. 11, the bearing load P is set within a range other than the range (P1 to P2) of the shaded part corresponding to the frequency range of the oscillation natural frequency “2 · f _o ” including the margin part. Just do it.

The bearing load P changes the bearing alignment A (here, the amount of movement when the center of the bearing (bearing inner ring 1) is moved vertically upward with respect to the center of the rotating shaft (journal 3)). It is possible to adjust by doing. FIG.
FIG. 13 is a diagram showing the relationship between the two by taking the bearing alignment A on the horizontal axis and the change in the bearing load P with respect to the change in the alignment A on the vertical axis. As shown in FIG. The relationship is proportional. Therefore, bearing loads P1 to P2 that cause multiple harmonic resonance
The bearing alignment A may be set within a range other than the bearing alignments A1 to A2 corresponding to.

To adjust the bearing alignment A, when the bearing inner ring 1 and the bearing outer ring 4 (bearings) are attached to the bearing stand 5, the distance vertically above the center of the rotary shaft of the journal 3 of the journal 3 is A1 to. Adjust the bolts 6 etc. so that they are within the range other than A2, and attach them.

When the pad type journal bearing of this embodiment is used as a bearing for a journal of a rotary shaft of a steam turbine or the like, the bearing base 5 undergoes thermal expansion, so that the bearing alignment A changes. Therefore, the bearing alignment A may be adjusted based on the empirically obtained change amount.

[0082] Thus, by adjusting the bearing load P, 1/2 times the operational rotational speed f _o rocking natural frequencies (1
/ 2 · f _o ), 1 × (f _o ), 2 × (2 · f _o ), etc. can be detuned with a margin, and harmonic resonance due to non-linear vibration of the pad is avoided. In addition, it is possible to provide a pad type journal bearing with improved stability.

When the pad type journal bearing of the present embodiment is used for the journal of the rotary shaft of a steam turbine or the like, a plurality of pad type journal bearings are required because the rotary shaft of the steam turbine is large. is there. In that case, the load P on each bearing is changed by relatively changing each alignment A1, A2, ... Of the plurality of bearings.
It suffices to adjust 1, P2, ...

(Fourth Embodiment) In the present embodiment, by adjusting the rotational rigidity k _θ of the pad, the rocking natural frequency of the pad with _respect to frequencies such as N times and 1 / N times the operating speed. The means for detuning with a margin will be described.

FIG. 13 is a front view of a pad showing a main part of the pad type journal bearing of the fourth embodiment. That is, the rotational rigidity k _θ of the pad 40 is normally the journal 3
It is formed by the oil film pressure in the pad 40 generated by the rotation of the pad 40. However, since the swing vibration of the pad 40 is a minute vibration, as shown in FIG. It can be replaced by the translational motion M1.
Further, as shown in FIG. 13, a recessed groove 41 is provided at one end 40b1 on the back side of the pad 40, and a bearing inner ring 42 is formed.
A concave groove 43 is also provided in a portion of the inner surface 42 of the pad facing the concave groove 41 at the rear end of the pad. A spring (coil spring) 44 for adjusting the oscillation natural frequency for restraining the vibration of the pad 40 is inserted between the groove 41 and the groove 43.

In the structure shown in FIG. 13, the rotational rigidity k _θ of the pad 40 can be expressed by the following equation.

[0087]

Equation 5] _{k θ = k θ b + k} θ s ...... (4) where, k _{theta b} is a rotating rigid formed by oil film pressure, k _{theta s} is the rotational stiffness of the spring 44.

That is, a spring system is constructed in which a spring 44 for adjusting the oscillation natural frequency is newly added in parallel to the rotational rigidity formed by the oil film pressure.

That is, the spring 44 for adjusting the oscillation natural frequency is
Adjust the elastic strength, etc. by changing the rotational stiffness k _{theta s} by adjusting the rotational stiffness k _theta overall pad, half the rocking natural frequencies of the operating rotation speed f _o of the pad 40 ( 1/2 · f _o), 1 times (f _o), it is possible to detuning with a margin with respect to the frequency of such as a two-fold (2 · f _o). Therefore, it is possible to avoid the harmonic resonance due to the non-linear vibration of the pad 40, and to provide the pad journal bearing with improved stability.

Further, FIG. 14 shows the rotational rigidity k of the pad 45.
_As another example for adjusting _θ , the leaf spring 4 is provided between the pad rear end 45b1 and the bearing inner ring inner surface 46a.
7 is provided. Further, FIG. 15 is a diagram for explaining another example for adjusting the rotational rigidity k _θ of the pad 48 as well. Back end 48 of pad 48
b1 is provided with a concave groove 49, and the inner surface 50 of the bearing inner ring 50 is
A concave groove 51 is also provided in a portion of the pad back surface end portion 48b1 facing the concave groove 49. And the groove 4
An air spring 52 for adjusting the oscillation natural frequency for restraining the vibration of the pad 48 is inserted between the groove 9 and the concave groove 51.

[0091] FIGS. 14 and 15 also adjusts the rotational stiffness of Figure 13 in the same manner as the plate spring 47 or the air springs 52 to adjust the rotational stiffness k _theta overall pad, rocking natural frequencies of the pads 40, 46 1/2 of the use rotational speed f _o (1/2 ·
f _o), 1-fold (f _o), it is possible to detuning with a margin relative to the frequency of such double (2 · f _o), to avoid harmonic resonance based on nonlinear vibration pad, It is possible to provide a pad type journal bearing having further improved stability without impairing the function of the conventional journal bearing.

As described in the description of FIG. 4,
Each pad has a spring 44 (47, 5 for adjusting the oscillation natural frequency).
2) is provided, and the rotational rigidity k _{θ s} of each spring 44 (47, 52) is changed according to the load applied to each pad to adjust the rotational rigidity k _{θ of the} entire pad, thereby further stabilizing the journal bearing. Can be provided.

In this embodiment, a coil spring, a leaf spring, an air spring, or the like is used as the member that imparts rotational rigidity to the pad, but the present invention is not limited to this.
Any member such as elastic means other than a spring may be used as long as it can impart rotational rigidity to the pad.

(Fifth Embodiment) Next, an embodiment of the second invention will be described. In the present embodiment, means for reducing the vibration amplitude by lowering the response sensitivity even if a harmonic resonance phenomenon occurs due to insufficient detuning of the oscillation natural frequency of the pad or the like will be described.

As described above with reference to FIG. 26, even if harmonic resonance occurs by damping the vibration system,
Since the vibration amplitude can be reduced, the damping mechanism is installed in the pad type journal bearing in this embodiment.

This damping mechanism only needs to provide damping to the system for the oscillating vibration of the pad that can be modeled by the equation (1).
As a means therefor, any mechanism may be used as long as it exerts a resistance force, which is proportional to the displacement speed, on the translational movement of the pad end portion shown in FIG.

Various damping mechanisms will be described below.

FIG. 16 shows, as an example of a damping mechanism, a squeeze effect generated in a thin oil film 62 between a back surface 60b of a pad 60 and an inner surface 61a of a bearing inner ring 61 (when two surfaces forming an oil film rapidly approach each other. In addition, a squeeze portion (squeeze damper) 63 utilizing a phenomenon in which a resistance force proportional to the approaching speed is generated)
It is installed at 1 and 60b2.

In a normal journal bearing, as shown in FIG. 17, the radius of curvature R _i of the back surface 60b of the pad 60 is set so that the pad 60 has a self-centering function.
It is designed to be smaller than the radius of curvature R _{o of} 1a. On the other hand, between the pad rear end portion 60b1 and the bearing inner ring inner surface 61a, the bearing lubricating oil is filled. However, since the curvatures of both are in the above-described relationship, the pad 60 makes a swinging motion and the pad Even if the back end 60b1 approaches the bearing inner ring inner surface 61a, the pad back end 60b1 and the bearing inner ring inner surface 61a
The gap between the pad 60 and the bearing inner ring 61 is widened toward the end, so that the lubricating oil S1 does not generate the squeeze effect, and the lubricating oil S1 passes from the gap C between the pad rear end 60b1 and the bearing inner ring inner surface 61a. It will be discharged. Therefore,
Squeeze effect occurs if you prevent the discharge of this lubricating oil as much as possible,
It becomes possible to give a large damping force to the system.

[0100] That is, as shown in FIG. 17, the region shown in the figure of the pad back end 60b1 "g", the squeeze portion 63 having a radius of curvature R _o of the same radius of curvature R _o of the inner bearing ring inner surface 61a It is provided. By this squeeze portion 63, the back surface of the squeeze portion 63 and the inner surface 61 of the bearing inner ring are substantially parallel to each other, and a squeeze oil film 62 is formed by the two parallel surfaces.

According to the above construction, the pad back surface end portion 60b1 does not become wider due to the squeeze portion 63, so that when the squeeze portion back surface 63 approaches the inner surface 61 of the bearing inner ring,
The amount of lubricating oil discharged can be minimized. As a result, a sufficient damping force is generated by the squeeze oil film 62 formed in the gap C between the back surface of the squeeze portion 63 and the bearing inner ring inner surface 61a, and the vibration of the pad 60 can be reduced.

FIG. 18 shows another example of the damping mechanism.
On the bearing inner ring inner surface 65a, end of the rear end portions 66b1, 66b2 of the pad 66 (pad open end) 66b _a
The discharge of the lubricating oil S1 that has flowed into the gap between the pad back end portions 66b1 and 66b2 and the bearing inner ring inner surface 65a described above is prevented at a position near the pad 65 so as not to affect the swing of the pad 65. Cough 67
Is provided.

According to FIG. 18, a sufficient damping force is generated by the squeeze effect of the lubricating oil S1 that is dammed in the gap between the pad back end portions 66b1 and 66b2 and the bearing inner ring inner surface 1a by the dam 67 and the pad 65 is formed. The vibration of can be reduced.

As shown in FIG. 17, in the example of the damping mechanism described above, a lubricating oil having a high clay temperature and a low temperature, which is different from the lubricating oil flowing into the bearing, is applied to the pad back end portions 66b.
An oil supply line 68 may be provided in the bearing inner ring 65 in a penetrating manner so that the oil supply port communicates with the gap in order to supply into the gap between the 1, 66b2 and the inner surface 65a of the bearing inner ring. In this case, the squeeze effect can be effectively and reliably generated by the lubricating oil supplied through the oil supply pump and the oil supply line 68 (not shown) into the gap.

That is, according to the present embodiment, even when a harmonic resonance phenomenon occurs, a damping force is applied to the oscillation vibration of the pad by using, for example, the squeeze effect of lubricating oil, so that the oscillation of the pad is oscillated. Vibration can be suppressed.
Therefore, it is possible to provide a pad-type journal bearing that is stable against vibration.

In this embodiment, the squeeze effect of the lubricating oil is used to provide a damping force to the oscillation vibration of the pad, but the present invention is not limited to this, and the pad end is not limited to this. Any structure may be used as long as it is a mechanism that gives the pad a resistance force proportional to the displacement speed of the translational movement of the part.

(Sixth Embodiment) Next, an embodiment of the third invention will be described. In the present embodiment, even if the pad resonates at the oscillation natural frequency, or even if it resonates at a harmonic, a means for actively controlling the oscillation of the pad so as not to adversely affect the rotation axis and damping. (Active control means) will be described.

FIG. 19 shows a sixth embodiment equipped with active control means.
It is a front view of a pad showing a main part of a pad type journal bearing of an embodiment. As described in the fourth embodiment, since the swing vibration of the pad is a minute vibration, it can be replaced with a translational motion along the radial direction of the rotation axis of the pad end as shown in FIG.

According to FIG. 19, in the pad type journal bearing, in order to detect the displacement amount of the translational motion of the pad 70, one end portion on the back surface side of the pad (in FIG. 19, it is the left end, which is hereinafter referred to as the left end portion). ) Displacement sensor 7 near 70b1
1 is installed. The displacement sensor 71 is embedded in, for example, the bearing inner ring 72 in the present embodiment. The pad type journal bearing is connected to the displacement sensor 71, and a converter 73 that converts the displacement amount detected by the displacement sensor 71 into an electric displacement signal, and the converter 7
3 and a frequency analyzer 74 having a filter function for decomposing the displacement signal converted by the converter 73 into a vibration frequency component targeted for active control.

Further, according to FIG. 19, the pad type journal bearing is embedded in the bearing inner ring 72, and the opening end thereof is the other end portion on the pad back surface side (in FIG. 19, it is the right end when facing,
A pipe 75 is provided so as to communicate with the gap between the right end portion 70b2 and the bearing inner ring inner surface 72a. This pipe 75 includes a valve 76 and a pipe 77.
It is connected to the boost pump 78 via. The valve 76 used in the present embodiment has the ability to open / close at high speed under the control of a valve controller described later, and can open / close at several hundred Hz, for example.

On the other hand, the control terminal of the valve 76 has a valve controller 7 for electrically controlling opening / closing of the valve 76.
9 is connected.

Further, a controller 80 is connected to the frequency analyzer 74 and the valve controller 79. The controller 80 calculates the opening / closing timing of the valve 76 for damping the vibration based on the vibration frequency component sent from the frequency analyzer 74, and sends the opening / closing timing information to the valve controller 79. ing. The valve controller 79 is configured to open / close the valve 76 according to the opening / closing timing information.

Next, the overall operation of this embodiment will be described.

When rocking vibration occurs in the pad 70, the displacement sensor 71 detects the amount of displacement of the pad left end portion 70b1 accompanying the rocking vibration. The displacement amount of the pad 70 is converted to a frequency component for active control via the converter 73 and the frequency analyzer 74, and then sent to the controller 80.

In the controller 80, the frequency analyzer 74
The opening / closing timing of the valve 76 for damping the vibration is calculated based on the vibration frequency component sent from the device, and the opening / closing timing information is sent to the valve controller 79. As a result, the valve 76 opens and closes under the control of the valve controller 79.

On the other hand, the valve 76 is constantly supplied with high-pressure oil from the booster pump 78 via the pipe 77. When the valve 76 is opened, the high-pressure oil S2 passes through the pipe 75 and the pad rear surface as shown in FIG. It is supplied to the gap between the right side end portion 70b2 and the inner surface 72a of the bearing inner ring and acts on the pad back surface 70b. As described above, since the valve 76 has the ability to open / close at high speed, it can open / close while sufficiently following the vibration that changes at high speed.

Since the high-pressure oil S2 acts on the pad right end rear surface 70b2, it is possible to prevent the pad right end rear surface 70b2 from approaching the bearing inner ring inner surface 72a and suppress the vibration of the pad.

On the other hand, when the pad right end rear surface 70b2 moves away from the bearing inner ring inner surface 72a, the back side left end portion 70b1 of the pad 70 conversely approaches the bearing inner ring 72a. Therefore, a mechanism provided on the pad right end portion ( Pipe 75 ',
A valve 76 ', a pipe 77', a booster pump 78 ', etc. are provided also on the left end side of the pad, and the control mechanism (displacement sensor 7) is provided.
1, converter 73, frequency analyzer 74, controller 8
0, the valve controller 79, etc.) to control the opening and closing of the valve 76 'to prevent the pad left end rear surface 70b1 from approaching the inner surface 72a of the bearing inner ring due to vibration, and suppress the vibration of the pad 70. it can. In other words, by providing the active control mechanism at both ends of the pad,
Effective pad vibration control is possible.

FIG. 20 is a front view of a pad showing a main part of a pad type journal bearing provided with other active control means.

In FIG. 20, the displacement sensor 71, the converter 7
The configurations of the frequency analyzer 74, the controller 80, and the valve controller 79 are the same as those in FIG. 19, and the same reference numerals are given and the description thereof is omitted. In the active control means shown in FIG. 21, the recess 91 provided on the back surface 90b2 on the right end portion side of the pad 90 and the recess 9 provided on the back surface 90b2 on the pad right end portion of the inner surface 92a of the bearing inner ring 92.
1 and the concave portion 9 provided in the portion facing 1
1 and a recess 93, and a bellows 94 inserted between the bellows 94 and the recess 93, and a pipe 95 connected to the bellows 94 for supplying high-pressure oil into the bellows 94. The pipe 95 is connected to a boost pump 98 via a (three-way) valve 96 which is a three-way valve and a pipe 97. Then, as in FIG. 20, a valve controller 79 that electrically controls opening / closing of the valve 96 is connected to the control terminal of the valve 96.

In the structure shown in FIG. 19, high-pressure oil is directly applied to the pad back surface to control the pad vibration, but in the structure shown in FIG. 20, the boost pump 98 to the valve 96 are used.
The high pressure oil is supplied into the bellows 84 by the opening operation of the pad, and the expansion of the bellows 94 causes the rear surface 90b1 of the right end portion of the pad.
Of the pad 90 to the inner surface 92a of the bearing inner ring.
The vibration of can be suppressed. In the configuration of FIG. 20, high pressure oil is supplied to the bellows 94 to supply the pad back surface 90.
If the bellows is filled with the working fluid (high-pressure oil) after the damping force is applied to b, the function of the pad itself may be impaired.
The high pressure oil filled in the bellows 94 can be discharged through one valve (oil relief valve) 99 of the above.

As described above, according to the active control means shown in FIG. 20, the repeated damping force following the vibration of the pad 90 can be directly applied to the back surface of the pad by the bellows 94. Active control for rocking vibration can be performed efficiently.

As in the case of FIG. 19, a mechanism (bellows 94 ', pipe 9) provided on the right end side of the pad is used.
5 ', valve 96', pipe 97 ', high pressure pump 98'
Etc.) may be provided on the left end side of the pad to perform active control at both ends of the pad.

As described above, as shown in FIGS. 19 and 20,
Even if the pad resonates at the oscillation natural frequency or if it resonates at the harmonic resonance, the oscillation vibration of the pad can be actively controlled (active control) without adversely affecting the rotation axis. It is possible to provide a pad-type journal bearing excellent in stability without impairing the functions of the type journal bearing (the function of supporting the journal by the oil film of the lubricating oil, the self-aligning function due to the change in the oil film pressure distribution).

Further, when the pad type journal bearing is applied to a rotary machine in which the bearing load changes over a wide range according to the operating condition, the rocking natural frequency of the pad changes as the load changes. Since there is a change, there are cases in which the above-described adjustment of the oscillation natural frequency of the pad is insufficient. However, as described in detail in the present embodiment, if the pad oscillation natural frequency is positively controlled, it is possible to sufficiently cope with the change in the pad oscillation natural frequency due to the change in the load.

In the pad-type journal bearings described in the first to sixth embodiments, the number of pads is not limited, and may be six, for example, as in the structure shown in FIG. 21. As shown in the form, it may be four. That is, the present invention can be effectively used for a pad type journal bearing configured by using one or more pads. Furthermore, in the pad-type journal bearings described in the first to sixth embodiments, the explanatory diagram in which the resultant force of the bearing load acts on the pad surface is used, but the present invention is not limited to this, and the bearing load It may be a pad type journal bearing in which the pads are arranged so that the resultant force of the above acts between the pads.

[0127]

As described above, in the pad type journal bearing, the pad causes a non-linear oscillating vibration (non-linear vibration) with respect to the exciting force received from the journal part, and the oscillating natural frequency of the pad causes the journal part to have a oscillating natural frequency. If the excitation frequency is equal to or close to N times and 1 / N times (N is a natural number equal to or larger than 1) the rotation speed of, the harmonic resonance will occur. This harmonic resonance develops into unstable vibration of the entire rotary shaft system, which hinders stable operation of a large rotating machine such as a steam turbine in which a pad type journal bearing is used.

However, according to the first aspect of the invention, as the detuning means, for example, the radius of curvature of the back surface of the pad or the radius of curvature of the inner peripheral surface of the bearing portion where the pad is in contact is adjusted to adjust the vibration of the pad. The frequency of the dynamic vibration can be detuned from the values of N times and 1 / N times (N is a natural number of 1 or more) times the rotation frequency of the journal portion with a predetermined margin, for example. Therefore, it is possible to avoid the above-mentioned harmonic resonance due to nonlinear vibration, and it is possible to provide a pad-type journal bearing that is vibrationally stable.

According to the second aspect of the invention, since the radius of curvature of the pad back end and the radius of curvature of the inner peripheral surface of the bearing are made substantially the same by the squeeze part as the damping mechanism, The squeeze effect is efficiently generated in the lubricating oil flowing into the gap between the bearing portion and the inner peripheral surface, and the squeeze effect can attenuate the vibration amplitude of the pad due to the harmonic resonance. As a result, it is possible to sufficiently suppress the above-mentioned harmonic resonance due to the non-linear vibration, and it is possible to provide a pad-type journal bearing that is vibrationally stable.

Further, according to the third invention, the active control means supplies the lubricating oil to the gap according to the vibration displacement of the pad, and the vibration of the pad due to the harmonic resonance causes the rotation of the journal portion. Active control is performed to a predetermined value that does not affect vibration. As a result, vibration associated with harmonic resonance can be suppressed, and a pad-type journal bearing that is vibrationally stable can be provided.

As described above, according to the present invention, it is possible to provide a pad type journal bearing that is vibrationally sufficiently stable. Therefore, it is unstable in a large rotating machine such as a steam turbine equipped with the pad type journal bearing according to the present invention. Vibration can be suppressed as much as possible, and stable operation for a long period of time becomes possible.

[Brief description of the drawings]

FIG. 1 is a front view showing a main part of a pad type journal bearing according to a first embodiment of the present invention.

FIG. 2 is a front view showing a main part of a pad type journal bearing according to a modified example of the first embodiment of the present invention.

FIG. 3 is a diagram showing a relationship between a pad clearance C _r and a rocking natural frequency and a detuning range of the rocking natural frequency.

FIG. 4 is a front view showing a main part of a pad type journal bearing according to a modified example of the first embodiment of the present invention.

FIG. 5 is a diagram showing a relationship between a pad mass and a rocking natural frequency of the pad and a detuning range of the rocking natural frequency.

FIG. 6 is a perspective view of a pad, which is a main part of a pad type journal bearing according to a second embodiment of the present invention, as viewed from the back side.

FIG. 7 is a perspective view of a pad, which is a main part of a pad type journal bearing according to a modified example of the second embodiment of the present invention, as viewed from the back side.

FIG. 8 is a perspective view of a pad, which is a main part of a pad type journal bearing according to a modified example of the second embodiment of the present invention, as viewed from the back side.

FIG. 9 is a front view showing a pad which is a main part of a pad type journal bearing according to a modified example of the second embodiment of the present invention.

FIG. 10 is a front view showing a main part of a pad type journal bearing according to a modified example of the second embodiment of the present invention.

FIG. 11 is a diagram showing a relationship between a bearing load and a rocking natural frequency and a detuning range of the rocking natural frequency.

FIG. 12 is a diagram showing a relationship between bearing alignment and bearing load and a range for ensuring detuning of oscillation natural frequency.

FIG. 13 is a front view showing a main part of a pad type journal bearing according to a third embodiment of the invention.

FIG. 14 is a front view showing a main part of a pad type journal bearing according to a modified example of the third embodiment of the present invention.

FIG. 15 is a front view showing a main part of a pad type journal bearing according to a modified example of the third embodiment of the present invention.

FIG. 16 is a front view showing a main part of a pad type journal bearing according to a fourth embodiment of the invention.

FIG. 17 is a front view for explaining the conventional discharge of lubricating oil from a gap.

FIG. 18 is a front view showing a main part of a pad type journal bearing according to a modified example of the fourth embodiment of the present invention.

FIG. 19 is a front view showing a main part of a pad type journal bearing according to a fifth embodiment of the invention.

FIG. 20 is a front view showing a main part of a pad type journal bearing according to a modified example of the fifth embodiment of the present invention.

FIG. 21 is a vertical sectional view showing the structure of a pad type journal bearing.

FIG. 22 is a diagram schematically showing the function of a pad type journal bearing.

FIG. 23 is a view for explaining a mechanism of rocking vibration of a pad in a pad type journal bearing.

FIG. 24 is a vibration chart showing an example of asynchronous unstable vibration generated in a conventional pad type journal bearing.

FIG. 25 is a diagram particularly showing a vibration amplitude of a pad as a response characteristic with respect to a frequency ratio of an exciting force for exciting the pad in the conventional pad type journal bearing.

FIG. 26 is a diagram showing a vibration amplitude as a response characteristic of the pad with respect to a frequency ratio of an exciting force for exciting the pad with the damping ratio as a parameter.

[Explanation of symbols]

1 Bearing Inner Ring 1a Bearing Inner Ring Inner Surface 2 Pad 2a Inner Surface 2b Rear Surface 3 Journal 4 Bearing Outer Ring 5 Bearing Stand 6 Bearing Mounting Bolt 7 Fixing Hole 8 Support Pin 9 Lubrication Line 10 Gap 11 Drain 21a Bearing Inner Ring Inner Surface 22 Pad 22a Pad Inner Surface 22b Pad back surface 23 Pad 23b Pad back surface 24a Bearing inner ring inner surface 24b Pad contact surface 25a to 25d Pad 26, 26A, 26B, 26C, 26D Pad 27 Horizontal hole 27A Vertical hole 28, 29 Groove 30 Reduction part 31, 42, 46 Bearing inner ring 31a, 42a , 46a Bearing inner ring inner surface 40, 45, 48 Pad 40b, 48b, 52a Pad back surface 48b1 41, 43, 49 Recessed groove 44 Spring 47 Leaf spring 48b1 Pad back end 52 Air spring 60, 66 pad 60b, 66b Pad back surface 61, 5 Bearing inner ring 61a, 65a Bearing inner ring inner surface 62 Oil film 63 Squeeze part 64 66b1, 66b2 Back end parts 67 Weir 70, 90 Pad 70b, 90b Pad back 70b1, 70b2, 90b1, 90b1 Pad back end 72, 71 Displacement sensor 92 bearing inner ring 72a, 92a bearing inner ring inner surface 73 converter 74 frequency analyzer 75, 95 pipe 76 valve 77, 97 pipe 78 booster pump 79 valve controller 80 controller 91, 93 recess 94 bellows 96 three-way valve 98 booster pump 99 oil Relief valves R _i , R _ia Pad back surface curvature radius R _o Bearing inner ring inner surface curvature radius R _oa Pad contact surface curvature radius C _rB Pad clearance m, m1, m2 Pad mass P, P1, P2 Bearing load A, A1, rotational stiffness M1 translation luck A2 bearing alignment k _theta pad rotational stiffness S1 lubricant rotational stiffness k _{theta s} spring formed by k _{theta b} oil film pressure

 ─────────────────────────────────────────────────── ─── Continuation of the front page (72) Inventor Toshio Hirano 4-4, 2 Suehiro-cho, Tsurumi-ku, Yokohama-shi, Kanagawa Toshiba Keihin Office

Claims (30)

[Claims] 1. A cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface, and the pad is swingable with the back surface in contact with the inner peripheral surface of the bearing portion. In a pad type journal bearing that is arranged on a pad type journal bearing that automatically adjusts the oil film pressure of the lubricating oil by swinging vibration of the pad to support the journal portion of the rotating shaft in a self-aligning manner, N times and 1 / N times the rotational frequency of the journal section (N is a natural number of 1 or more)
A pad type journal bearing, characterized in that it is provided with a detuning means for detuning from the value of. 2. The detuning means sets the frequency of rocking vibration of the pad to N times and 1 / the rotational frequency of the journal.
The pad-type journal bearing according to claim 1, wherein the pad-type journal bearing is detuned with a predetermined margin from the value of N times. 3. As the detuning means, the backside of the pad is detuned with a predetermined margin from the values of the oscillation frequency of the pad, which is N times and 1 / N times the rotational frequency. The pad-type journal bearing according to claim 2, wherein the pad-type journal bearing is formed with a radius of curvature that provides a specified frequency. 4. The plurality of pads are arranged concentrically with each pad in contact with the inner peripheral surface of the bearing portion, and the radius of curvature of the back surface of each pad depends on each pad. The method according to claim 3, wherein each of them is set according to a bearing load. 5. As the detuning means, at a position on the inner peripheral surface of the bearing portion where the pad contacts, the vibration frequency of the rocking vibration of the pad is predetermined from a value which is N times or 1 / N times the rotational frequency. 3. The pad type journal bearing according to claim 2, further comprising a pad contact surface formed with a radius of curvature that provides a detuned frequency with a margin of 1. 6. The pad comprises a plurality of pads, the pads being arranged concentrically in a state of being in contact with the inner peripheral surface of the bearing portion, and each pad being formed at a position in contact with the pad. The pad journal bearing according to claim 5, wherein the radius of curvature of the surface is set according to the bearing load applied to each pad. 7. The detuning means detunes the mass of the pad with a predetermined margin from the values of the oscillation frequency of the pad, which is N times and 1 / N times the rotational frequency. The pad-type journal bearing according to claim 2, wherein the pad-type journal bearing is set to have a specified frequency. 8. The pad-type journal according to claim 7, wherein the pad is provided with a required number of holes having a required shape and a required length along the axial direction of the journal to adjust the mass of the pad. bearing. 9. The pad-type journal according to claim 7, wherein the pad is provided with a required number of holes having a required shape and a required length along the radial direction of the journal to adjust the mass of the pad. bearing. 10. The pad type journal bearing according to claim 7, wherein a required number of grooves having a required shape are provided at the rear end portion of the pad to adjust the mass of the pad. 11. The pad type journal bearing according to claim 7, wherein the pad is provided with notches having a required shape and a required length at both end portions on the back surface side to adjust the mass of the pad. 12. The mass of the pad is adjusted by setting a central angle of the pad to a required angle.
Pad type journal bearing described. 13. The pad comprises a plurality of pads, the pads being concentrically arranged in contact with the inner peripheral surface of the bearing portion, and the mass of each pad is the bearing load applied to each pad. 13. The pad type journal bearing according to claim 7, wherein the pad type journal bearing is set according to each item. 14. As the detuning means, a bearing load applied to the pad is obtained by taking a predetermined margin from the values of the oscillation vibration of the pad being N times and 1 / N times the rotational frequency. The pad type journal bearing according to claim 2, wherein the pad type journal bearing is set to have a detuned frequency. 15. The bearing load is set by having a bearing stand for fixedly supporting the bearing portion, and changing the fixed support position of the bearing portion to adjust the bearing alignment of the bearing portion to a required value. Item 14. The pad type journal bearing according to Item 14. 16. As the detuning means, a rotational rigidity imparting means for imparting a predetermined rotational rigidity to the pad is inserted between the pad back surface and the inner peripheral surface of the bearing portion, and the rotational rigidity imparting means is provided. The vibration frequency of the rocking vibration of the pad is multiplied by N and 1 / N of the rotation frequency according to the rotational rigidity given by
The pad-type journal bearing according to claim 2, wherein the frequency is set to a detuned frequency with a predetermined margin taken from the doubled value. 17. The pad type journal bearing according to claim 16, wherein the rotational rigidity imparting means is an elastic member having a predetermined elastic force. 18. The pad type journal bearing according to claim 17, wherein the elastic member is a coil spring. 19. A first groove is provided on the back surface of the pad,
The second groove is provided at a position of the bearing portion facing the first groove, and the coil spring is inserted between the first groove and the second groove. Pad type journal bearing described. 20. The elastic member is a leaf spring.
7. Pad type journal bearing according to 7. 21. The pad type journal bearing according to claim 20, wherein the leaf spring is provided on the back surface of the pad. 22. The pad type journal bearing according to claim 17, wherein the elastic member is an air spring. 23. The plurality of pads are arranged concentrically with each pad being in contact with the inner peripheral surface of the bearing portion, and the rotational rigidity given to each pad is different from each other. The pad type journal bearing according to any one of claims 16 to 22, which is set according to the bearing load. 24. A cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface, and the pad is swingable with the back surface in contact with the inner peripheral surface of the bearing portion. In a pad type journal bearing which is arranged on the pad and which supports the journal portion of the rotating shaft in a self-aligning manner by changing the lubricating oil film pressure by the oscillation vibration of the pad, when a harmonic resonance phenomenon occurs in the pad, , A pad-type journal bearing comprising a damping mechanism for damping the vibration amplitude of the pad due to the harmonic resonance. 25. The damping mechanism includes a squeeze portion provided at an end portion on the back surface side of the pad so that the radius of curvature of the end portion of the pad back surface is substantially the same as the radius of curvature of the inner peripheral surface of the bearing portion. The pad type journal bearing according to claim 24, which has. 26. The damping mechanism prevents movement of the lubricating oil that has flowed into a gap between the pad rear end portion and the bearing inner ring inner surface in the vicinity of the pad rear end portion and in the swing of the pad. 25. The pad type journal bearing according to claim 24, further comprising a weir portion provided at a position separated from the pad so as not to affect. 27. An oil supply line communicating with a gap between the pad rear end portion and the bearing inner ring inner surface is provided, and a clay different from the lubricating oil having a high temperature and a low temperature is provided in the gap via the oil supply line. The lubricating oil of No. 2 is supplied.
The pad type journal bearing according to 5 or 26. 28. A cylindrical bearing portion and at least one arc-shaped pad having a spherical back surface are provided, and the pad is swingable with the back surface in contact with the inner peripheral surface of the bearing portion. In a pad type journal bearing which is arranged on the pad and which supports the journal portion of the rotating shaft in a self-aligning manner by changing the lubricating oil film pressure by the oscillation vibration of the pad, when a harmonic resonance phenomenon occurs in the pad, A pad type journal bearing comprising: active control means for actively controlling the vibration of the pad due to the harmonic resonance to a predetermined value that does not affect the rotational vibration of the journal portion. 29. The active control means is provided at least near one end of the back surface of the pad, and detects the vibration displacement of the pad, and the active control target is based on the displacement data detected by the detection means. Generating means for generating vibration information, an oil supply line communicating with at least a gap between the other end of the back surface of the pad and the inner peripheral surface of the bearing portion, a supply means for supplying lubricating oil to the oil supply line, and the oil supply line 29. The pad type journal bearing according to claim 28, further comprising: a valve inserted in the middle of the valve to control opening / closing of the oil supply line; and valve opening / closing timing control means for controlling opening / closing timing of the valve according to the vibration information. 30. The active control means is provided at least in the vicinity of one end of the back surface of the pad, and detects the vibration displacement of the pad, and the active control target is based on the displacement data detected by the detection means. Generating means for generating vibration information, a bellows inserted between at least the other end of the pad back surface and the inner peripheral surface of the bearing, an oil supply line connected to the bellows, and the oil supply line. Supply means for supplying lubricating oil, a three-way valve inserted in the middle of the oil supply line to control opening / closing of the oil supply line, and valve opening / closing timing control means for controlling opening / closing timing of the three-way valve according to the vibration information. 29. The pad type journal bearing according to claim 28, further comprising: