JPS59212513A - Bearing apparatus - Google Patents

Abstract

PURPOSE:To control the dangerous speed to a constant value by continuously forming a pressing mechanism for pressing a bearing towards the outer peripheral surface side of a rotary shaft, onto the outer peripheral part in the radial direction of the bearing. CONSTITUTION:An annular hydraulic chamber 2 is formed continuously onto the outer peripheral part in the radial direction of a bearing 1 for supporting a rotary shaft S and a pressing mechanism PM for pressing the bearing 1 towards the outer peripheral surface side of the rotary shaft S is formed. The oil discharged from the pump P of a hydraulic pressure feeding apparatus 5 is supplied into the hydraulic chamber 2 through an oil passage 4. A relief valve 3 for controlling the amount of oil returned to a reservoir R is installed in parallel to the pump P. Thus, the gap between the shaft S and the bearing 1 can be adjusted by controlling the pressure P acting into the hydraulic chamber 2 can be controlled.

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a bearing device for supporting a shaft rotating at high speed, such as a winding and unwinding reel of a rolling equipment.

In the conventional bearing device of this type, Fig. 1 (j), (b)
As shown in the figure, in most cases, the rotating shaft S is rotated while being supported at one or two points by a bearing L.

When the shaft S is rotated at high speed, the number of revolutions that can be rotated is determined by the method of supporting the shaft S, the diameter of the shaft S, and the length l. In other words, at a speed higher than that, the number of shafts S matches and resonates, and the shaft S
may be damaged. The rotational speed at this time is called the critical speed.

This critical speed depends on the support method (one-point support or two-point support), the diameter and length of the shaft, and the shaft structure as shown in Figure 2 (a), as shown in Figures 1 (a) and (b). A double shaft structure with a sleeve Sa interposed between the solid shaft S and the bearing 1, or a double shaft structure as shown in FIG.
) As shown in ), it also differs depending on whether it is a single solid shaft structure in which a solid shaft S is supported by a bearing 1. Also, the higher the natural frequency, the higher the critical speed.

Next, experimental results regarding the natural frequency characteristics of the single solid shaft structure and double shaft structure described above will be explained. Figure 3 (b
) is the result.

This is a bearing device (
(almost the same structure in the case of the double shaft structure), the motor M
In this case, the shaft S is rotated at a rotation speed N=1000 rpm, but when the shaft S is supported by the bearing 1 and rotated, the support part (
Heat at this neck (bearing temperature)
The relationship between To and the natural frequency is shown for each rising temperature of bearing 1 in the case of a single solid shaft structure (solid line) and in the case of a double shaft structure (dotted line).
The characteristic diagram shown in FIG. 3(b) summarizes these characteristics.

As can be seen from FIG. 3(b), in the case of the double-shaft structure as shown in FIG. 2(a), the natural frequency is approximately constant regardless of the temperature rise of the bearing l. [Refer to the characteristics indicated by the dotted line in Figure 3 (b)] On the other hand, if the −-shaped solid shaft structure is used as shown in Figure 2 (b), the natural frequency increases as the temperature of the bearing 1 increases. There is. [Refer to the characteristics indicated by the solid line in FIG. 3(b)] As a result of investigating the cause of this, the following was found.

First, in the double-shaft structure, as shown in FIG. 4(a), there is a very small gap δ1 between the bearing 1 and the sleeve SC.

Furthermore, there is also a gap δ2 between the sleeve SC and the shaft S, which is larger than δ1.

By the way, if the temperature of the bearing 1 rises during rotation, the temperature of that portion of the shaft will also rise. As a result, the sleeve Sa expands due to heat, and the bearing 1 also expands.

The shape 9 of the sleeve SC is better than the expansion of the bearing 1.
Due to its large size, the sleeve SC expands and gradually reduces the gap δ1 between the bearings 1, eventually becoming δ, -0, and finally a compressive force is generated between the bearing 1 and the sleeve Sa.

The shaft S also expands due to temperature rise, but the sleeve SC
The clearance δ2 between the sleeve SC and the bearing 1 is much larger than the clearance δ2 between the sleeve SC and the bearing 1.
It was found that the gap only decreases, but the gap always exists.

Next, in the case of a - type solid shaft structure, there is a gap between the bearing 1 and the shaft S as shown in FIG. 4(b), and between the bearing 1 and the sleeve SC in FIG. 4(a). There is a gap δ3 of the same amount. In this case as well, as the temperature rises, the gap δ3 narrows, and finally, like the gap δ1 in FIG. 4(a), the gap δ3 disappears, and it was found that the amount of compression begins to act during this time.

From this result, it can be seen that in the case of the - type solid shaft structure, as the temperature rises, the gap δ3 between the bearing 1 and the shaft S disappears, and the spring constant K increases, resulting in an increase in the natural frequency. Conversely, in the case of a double shaft structure, the clearance δ1 between the sleeve SC and the bearing 1 disappears, but the clearance δ2 remains between the shaft S and the sleeve Sa, so the spring constant does not increase. First, it turned out that it has nothing to do with the natural frequency.

Here, the spring constant is a constant determined from the correlation between the deflection ΔS of the object and the reaction force p thereto, and is given by the angle K as shown in FIG. FIG. 5 shows the change in spring constant that would be obtained when the clearance between the bearing 1 and the shaft S was changed from large to medium to small. In other words, the smaller the gap, the larger the spring constant.

Figure 6(b) shows the relationship between the spring constant and critical speed using a bearing device with dimensions as shown in Figure 6(a) (almost the same structure in the case of double shaft structure). These are calculation results comparing the case of a single solid shaft structure and the case of a double shaft structure.

As can be seen from this calculation result, the greater the spring constant K, the greater the critical speed. In other words, the larger the spring constant K is, the higher the rotation speed becomes possible.

An increase in the critical speed means an increase in the natural frequency.

From the above results, if there is a support means that increases the spring constant K of the rotating shaft system, and a support means that can adjust this spring constant K to a constant value, it is possible to always maintain a constant critical speed and stabilize and improve the safety of operations. I can keep it.

The present invention was created based on the above-mentioned knowledge obtained by the inventor, and is a bearing that enables the spring constant of the rotating shaft system to be freely adjusted and the critical speed of the rotating shaft to be freely controlled. The purpose is to provide equipment.

For this reason, the bearing device of the present invention is characterized in that a pressing mechanism for pressing the bearing toward the outer peripheral surface of the rotating shaft is connected to the radially outer peripheral portion of the bearing that supports the rotating shaft.

Hereinafter, a bearing device as an embodiment of the present invention will be described with reference to the drawings. FIG. 7 is a schematic diagram showing the schematic configuration thereof, and FIG. 8 is a schematic diagram showing the schematic configuration of a modification thereof, In FIGS. 7 and 8, the same reference numerals as in FIGS. 1 to 6 indicate substantially similar parts.

As shown in FIG. 7, a bearing 1 is provided to support the high-speed rotating shaft S, and the bearing 1 is pushed toward the outer circumferential surface of the rotating shaft S on the radial outer circumference of the bearing 1. Annular hydraulic chambers 2 are connected to form a pressurizing mechanism PM.

Note that this hydraulic chamber 2 is attached to a fixed part (not shown), so that it can support the reaction force of the pressure p acting on the hydraulic chamber 2.

Further, a hydraulic pressure supply device 5 is connected to the hydraulic chamber 2 via an oil passage 4.
Discharge oil from the pump P is supplied.

In addition, an IJ-F valve 3 is provided in parallel with the pump P to control the amount of oil returned to the reservoir/<R. The clearance between the bearing 1 and the bearing 1 is adjusted so that a constant spring constant K is always obtained.

Since the spring constant K of the rotating shaft system can be adjusted to a constant value in this way, it is possible to prevent the critical speed and even the natural frequency from changing due to bearing temperature, contributing to safer and more stable operations. .

In addition, by adjusting the IJ'J valve 3, the value of the pressure p acting on the hydraulic chamber 2 can be changed, and the screw constant K of the rotating shaft system can be arbitrarily adjusted. The critical speed can be changed as appropriate depending on the situation, and as a result, the operating conditions can be expanded.

Note that the symbol M' in FIG. 7 indicates a pump driving motor.

By the way, as shown in FIG. 8, as the pressurizing mechanism PM,
By connecting the pair of wedges 6 and 7 to the outer circumference in the radial direction of the bearing 1 and by providing a mechanism for driving one of the wedges 7 in the direction of the arrow with a drive device 8 such as a hydraulic cylinder, the bearing 1 can be connected to the rotating shaft. The force p pressing toward the outer circumferential surface may be controlled. Note that the wedge 7 is restricted from moving outward.

With this configuration, the bearing 1 is tightened by a wedge effect,
The clearance and pressure between the bearing 1 and the shaft S can be adjusted so that a constant number constant K is always obtained, or it can be adjusted to an arbitrary number constant. In this case as well, it is possible to contribute to safer and more stable operations, and also to expand the range of operating conditions.

In addition to applying the pressure mechanism PM shown in Figs. 7 and 8 to a bearing device with a single solid shaft structure such as a double series, the pressure mechanism PM shown in Figs. 2(a) and 4(a) The present invention can also be applied to a bearing device with a dual shaft structure such as this.

As described in detail above, according to the bearing device of the present invention, a pressurizing mechanism that presses the bearing toward the outer peripheral surface of the rotating shaft is connected to the radially outer peripheral portion of the bearing that supports the rotating shaft. With a simple configuration that is installed in In addition, there are other benefits that can result in an expanded range of operating conditions.

[Brief explanation of the drawing]

Figures 1 to 6 show conventional bearing devices, and Figure 1 (a
), (b) are schematic diagrams showing different rotating shaft support means, Figures 2 (a) and (b) are schematic diagrams showing bearing devices with different structures, and Figure 3 (a) is a schematic diagram showing a bearing device with a different structure. A schematic diagram showing the approximate dimensions of various parts of the device used to determine the frequency-bearing humidity characteristic, Figure 3 (b) is a natural frequency-bearing temperature characteristic diagram, and Figures 4 (a) and (b) are A schematic diagram showing these structures in detail corresponds to Figures 2(a) and (b), respectively, and Figure 5 is a reaction force-deflection characteristic diagram to explain the method for determining the spring constant of the rotating shaft system. , Figure 6 (
a) is a schematic diagram showing the approximate dimensions of various parts of the device used to determine the critical speed-spring constant characteristic, Figure 6(b) is a diagram of the critical speed-spring constant characteristic, and Figures 7 and 8 are from this book. This shows a bearing device as an embodiment of the invention, and FIG. 7 is a schematic diagram showing the schematic structure thereof, and FIG. 8 is a schematic diagram showing the schematic structure of a modification thereof. 1. Bearing, 2. Hydraulic chamber, 3. IJ IJ valve,
4. Oil path, 5. Hydraulic supply device, 6, 7. Wedge,
8...Drive device, M'...Motor, P...Pump, R...
・Reservoir, S... Rotating shaft, PM... Pressure mechanism, SC...
·sleeve. Sub-Agent Patent Attorney Yoshiatsu Iinuma Figure 1 Figure 2 Figure 3 (0) (b) Axis - No. 7 □! ,;j JL'Tθ(0C) Fig. 4 (a) (b) Fig. 5 Tawa l, ·△S Fig. 6 (0) (b) Spring/T W Father K (kg/cm) No. 7 Figure 8 Continued from page 1 0 Author: Hikotabe Igaya 4-6-22 Kannon-Shinmachi, Nishi-ku, Hiroshima City Mitsubishi Heavy Industries, Ltd. Hiroshima Research Center 0 Author: Michinori Yamamoto 4-6-6 Kannon-Shinmachi, Nishi-ku, Hiroshima City No. 22 inside Mitsubishi Heavy Industries, Ltd. Hiroshima Shipyard

Claims (1)

[Claims] A bearing device comprising a pressure mechanism connected to a radially outer circumferential portion of a bearing that supports a rotating shaft and presses the bearing toward the outer circumferential surface of the rotating shaft.