JPS60146961A - Static double mechanical seal - Google Patents

Abstract

PURPOSE:To maintain good sealing performance even under variable pressure operation, by fitting an inner ring thinner than the outer ring while having diameter same with the seal diameter of seat ring receiver slidably and forming a retainer. CONSTITUTION:A retainer 11 is constructed by fitting an outer ring 11a over an inner ring 11b slidably in axial direction. Both faces of outer ring 11a are maintained approximately perfect planar where the inner ring 11b is thinner than the outer ring 11a while the outer diameter is made same with the seal diameter D of seat ring receiver 34b. Consequently, all forces to be applied onto the outer ring 11a will be the combination of pressure of enclosed liquid and pressing force of spring 37 where they are balanced perfectly to apply no moment onto the opposite side faces of outer ring 11a.

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a compact and economical stationary dual mechanical seal that can exhibit good sealing effects against pressure changes.

In stationary double-acting mechanical seals (hereinafter simply referred to as double-acting seals) for sealing rotary shafts built into various chemical industrial equipment, as the pressure of the fluid to be sealed increases, the amount of strain in the fluid-side seal increases. As a result, the problem is that the contact condition of the sealed end surface is poor and the sealing performance is deteriorated. That is, the configuration of this type of typical double seal is the third one.
The following explanation will be given based on FIG. 1 (a schematic vertical cross-sectional view of the main part). In FIG. 1, 31 is a rotating shaft, and 32 is a seal box for the rotating shaft 31. Also, 33a and 33b are seat rings disposed within the seal box 32 at predetermined intervals. In the arrangement, a plurality of pins 36 and barbs 37 are alternately inserted into the seat ring receivers 34a and 34b fixed at both ends of the seal box 32 along the circumferential direction. There is. The pins 36 are mounted in each pin hole to form a loose pin connection. That is, the seat ring 33a.

33b, respectively) IJ receivers 34a, 34.
It is attached so that it can be moved slightly forward and backward and up and down relative to b. and seat ring 33a.

33b is in contact with driven rings 40a and 40b, which will be described later, to form a double-sealed end surface.

The driven rings 40a and 40b are both integrally attached to the rotating shaft 31 so that they can rotate together with the rotating shaft 31, and their mounting structure is as follows. That is, both sides of the retainer 41 are formed to be completely flat, and the opposite end surfaces of the driven rings 40a and 40b are brought into contact with the two halves through rotation pins not shown in the drawings, respectively. Driven ring 40a + 40
b is an upper driven ring presser 44 and a lower driven ring presser 4
3, seven bodies are arranged on the left and right sides of the retainer 41.

Further, both side surfaces of the retainer 41 are lapped in advance and finished to an ultra-precise flat surface (surface roughness of approximately μ=0.0003 mm, flatness of approximately 0.7 μm or less). It goes without saying that the opposing end faces and sealed end faces of the driven rings 40a and 40b are lapped. In this way, the retainer 41 having the driven rings 40a+40b on the left and right sides is further mounted and positioned on the stepped portion 31a of the rotating shaft 31, and is fixed to the rotating shaft 31 by the sleeve 45 and the screw 46.

Also, 47a+47'a is high pressure cooling fluid (hereinafter referred to as sealing fluid)
The inlet 47b is the outlet for the sealing liquid, and the sealing liquid is sealed in the space 48 in the seal box 32, and a pressure slightly higher than that of the fluid to be sealed is applied to minimize the occurrence of thermal distortion at the sealed end surface. The seat ring 33a,
33b are sufficiently pressed against the driven rings 40a and 40b, respectively, to stabilize the sealed end surface.

Since the space 49 is loaded with the fluid to be sealed at a predetermined pressure, and the space 50 is loaded with atmospheric pressure, an O-ring is placed at an appropriate location for the purpose of isolating them from each other other than at the sealed end face. . In particular, 51, 52, and 54 are O-rings arranged in consideration of the seal between the fluid to be sealed and the sealing liquid), and 55, 56, and 58 are the O-rings that are arranged to provide a seal between the sealing liquid and the atmospheric pressure. This is an O-ring placed with sealing in mind.

With the above-described configuration, consideration has been given to ensuring sealing performance at the double-sealed end face.

By the way, as already mentioned, the problem with such high-pressure compound seals is that as the pressure of the fluid to be sealed increases as the operating pressure of the device increases, the sealing performance deteriorates rapidly. It is. This is shown in Figure 3 (b
) As shown in FIG. 3(a), which is an enlarged diagram for explaining the main part operation, when the pressure of the fluid to be sealed in the space 49 increases, the retainer 4
This is due to the fact that the thrust force (axial force as shown by the arrow in the figure) acting on the retainer 41 is too large to be ignored, and the amount of strain on the retainer 41 also increases sharply. The driven ring 40b, which is held by the driven ring 40b, also suffers from almost the same amount of distortion with the same distribution, and the gas-side sealed end face becomes the so-called open-face state as shown in the figure, resulting in a rapid increase in sealed fluid leakage. It is.

Therefore, as a means for solving such inconveniences, the present inventors made the retainer side end surfaces of the driven ring on the side of the fluid to be sealed in advance conform to the shape of the sides of the fluid to be sealed of the retainer under pressure of the fluid to be sealed. By forming and arranging them in a similar manner, we succeeded in developing a double seal that could guarantee long-term stability of sealing performance without making the structure of the mechanical seal part particularly large. I did it.
issue). However, it has become clear that a new problem arises when a device to which the dual seal of the earlier application is applied is operated in a so-called patch manner. In other words, the retainer side end face of the driven ring and the retainer are sealed for 9 hours from the start of operation to the specified pressurization operation (while the pressure is rising) and from the specified pressurization operation to the stop of the operation (while the pressure is falling). Since the sides of the end faces cannot face each other correctly, the sealed end faces are opened on both sides, which causes multiple sealed fluid leaks, and especially when the above-mentioned patch operation is performed at high pressure, the pressure fluctuations are large, so the leakage occurs. It became clear that problems with increasing volumes would arise. As a countermeasure to this problem, it may be possible to suppress the sealing fluid leakage within an allowable range by increasing the rigidity of the retainer and driven ring of the multiple seal, which was proposed by the present inventors in the earlier application. However, the size of the device is large and the material is expensive, so that it cannot meet the demands for compactness and reduction in manufacturing costs.

The present invention has been made in view of the above circumstances, and its purpose is to provide a compact and economical double seal that can maintain good sealing performance even during pressure fluctuation operation under patch operation. be.

However, the compound seal of the first invention that achieves the above purpose has an outer ring having both side surfaces formed into almost perfect planes, a thickness smaller than that of the outer ring, and an outer diameter smaller than that of the outer ring. The gist is that the retainer is constructed by fitting an inner ring having the same size as the seal diameter of the seat ring receiver so as to be slidable in the axial direction.

The configuration and operation and effects of the first invention will be explained below with reference to the drawings of the embodiment, but the description will focus on the characteristic configuration of the present invention, avoiding redundant explanation of the two basic configurations that are the same as the conventional example. do.

FIG. 1(a) is a diagram corresponding to the main part explanatory diagram of FIG. 3(b), and shows the set state of the double seal before pressurizing operation of the chemical reaction vessel. The retainer 11 can be said to have a configuration in which a conventional retainer (41 in FIG. 3) is divided into two in the circumferential direction. That is, the retainer 11 is formed by pushing the outer ring 11a onto the inner ring 11b so that it can slide in the axial direction, and both sides of the outer ring 11a maintain an almost perfect outer plane (precision finished surface similar to that of conventional retainers). Inner ring 11b is outer ring 11a
It is configured such that the thickness is small, and its outer diameter is the same as the seal diameter Dφmm of the seat ring receiver 34b. As a result, the force acting on the outer ring 11a of the retainer 11 (sealing fluid pressure P
′+pressing force S) of the spring 37, the two acting forces are perfectly balanced and no moment is applied to either side of the outer ring 11a. On the other hand, the inner ring 11b in the retainer 11
The pressure P of the sealing fluid acts on the end face of the sealing fluid, and its magnitude F is F=4X (D2'-d") XP (
KfIf/lllIn2) [where d is the inner diameter (+r+m,) of the inner ring 11b, and D is the inner diameter].

Due to this acting force, the inner ring 11b is distorted as shown in FIG. 1(b), but the inner ring 11b is fixed by the sleeve 45, and the thickness of the inner ring 11b is made appropriately smaller than the thickness of the outer ring 11a. Therefore, both side surfaces of the inner ring 11b, which is in a distorted state, can be prevented from coming into contact with the opposing end surfaces of the driven rings 40a and 40b.

Therefore, the outer ring 11a, which occupies the main part of the retainer 11, is not affected by the pressure P of the fluid to be sealed, so even during pressure fluctuation operation under patch operation, the vertical flatness of the sealed end face is maintained sufficiently uniform, ensuring good sealing performance. If you maintain it, you will be able to do it.

However, even with such a double seal according to the first invention, when batch operation is performed under pressure conditions of about 100 Kgf/c+f or more, the sealing performance cannot necessarily be said to be good, and especially when batch operation is performed at around 200 Kgf/c+f It was confirmed that when used to seal the rotating shaft of a polymerization reaction vessel, there is a greater concern that the sealing performance will deteriorate.

In order to eliminate such concerns, the present inventors closely observed the behavior of the multiple seal according to the first invention, and the following knowledge and solution guidelines were obtained.

That is, to explain based on FIG. 1(b), since the seat rings 33a and 33b are generally made of a super hard material such as WC, compared to the driven rings 40a and 40b made of a soft material such as carbon, It can be said that its rigidity is extremely high. Therefore, the pressure of about 100Kgf/7 or less is applied to the driven ring 40a.

Even if it acts on 40b, the retainer or driven ring 40a.

The problem of distortion generation as in 40b was almost negligible. Therefore, under such pressure conditions, the retainer 41
Good sealing performance should be maintained as long as distortion does not occur in the driven rings 40a and 40b, and as a result of conducting studies based on this policy, the first invention was completed. is the same as mentioned above. However, the sealing fluid pressure p
When / is about 100 bf or more, especially around 200 K7f%, the moment in the direction of the arrow acting on the retainer end portions A and B of each seat ring 33a and 33b cannot be ignored, and the end portions A and B It was found that some distortion occurred. As a countermeasure for this, the seat ring 33a
Although it is possible to increase the size of +33b itself, this would result in a larger seal structure and lead to an increase in manufacturing costs. Therefore, in order to prevent the above moment from occurring at the retainer protruding ends A and H without making the seat rings 33a and 33b particularly large, the ends A and B are not affected by pressure, that is, they are not affected by pressure. The guideline was that the structure should be balanced.

The second invention was made as a result of further cross-sections based on the above guideline, and it was found that the
Good sea 7.1 to 8. even during pressure fluctuation operation of chemical equipment that is batch operated at around 0 hf/c♂. , o,
3゜□63, 2.28, ah. , ' is intended to provide a double seal.

However, such a double seal of the second invention is a seal having an outer ring having both side surfaces formed into a nearly perfect plane, a thickness smaller than that of the outer ring, and an outer diameter of which is a seat ring receiver. The retainer is constructed by fitting an inner ring having the same size as the diameter so that it can slide in the axial direction, and further dividing each seat ring in the direction perpendicular to the axis and loosely connecting them with pins to abut each other. The gist is that the seal member is arranged so that the outer diameter of the seal is the same as the outer diameter of the sealed end surface.

The configuration and effects of the second invention will be explained below with reference to the drawings of the embodiments, but the description will also focus on the characteristic configuration of the second invention.

FIG. 2 is a diagram corresponding to the sectional explanatory view of the main part of FIG. 1(a), and shows the set state of the multiple seal before pressurizing the polymerization reaction vessel. The seat rings 21a+21b are both divided in the direction perpendicular to the axis and have sliding parts 21aI+21.
b, and supporting portions 21a2°21b2 are formed,
The structure of each pair of mb fitting part and support part is that both are QIJ
The rings 23a and 23b are held together and abutted at the bottom (while being sealed with an O-ring), and the bottles 22a and 22b are held at the top.
It is loosely coupled through the 0 ring, and further above the 0 ring'
23a.

The clamping position of 23b is configured such that the seal diameter d, φmm, is the same as the outer diameter of the sealed end surface. As a result, each seat ring 21a, 21b receives pressure (sealing fluid pressure P') equally from both sides, so that the components in the various arrow directions shown in FIG. 1(b) are It doesn't happen at all. Even if each seat ring 21alt21bl is slightly distorted in the radial direction, this does not cause any problem since it does not affect the vertical flatness of the sealed end face.

In this way, in addition to the basic effect of the pressure balance method by dividing the retainer 11 in the circumferential direction (the effect of preventing distortion from occurring in the driven rings 40a, 40b and maintaining the vertical flatness of the sealed end face), each seat ring 21a, The basic effect of the pressure balance method (the effect of maintaining the vertical flatness of the sealed end face by preventing the generation of a moment in the sliding portion 21 al l 2 l b+) can be enjoyed by dividing the portion 21b in the direction perpendicular to the axis. Therefore, as mentioned above, 100
Kyf/crr? More than 200 Kyt/c
Good sealing performance can be maintained even during pressure fluctuation operation of a polymerization reaction vessel which is batch operated at around m2.

It should be noted that each of the embodiments of the first invention and the second invention described above are representative examples and do not limit the present invention.

All changes in design such as shape and size belong to the technical scope of the present invention.

Furthermore, the mechanical seal of the present invention is not limited to application to the rotating shafts of chemical equipment as described above, but can also be applied to the rotating shafts of all industrial machinery, automobiles, aircraft turbines, etc. It can also be applied to aqueous solutions, lubricating oils, gases, etc.

The present invention is constructed as described above, but the point is that the vertical flatness of the sealed end face can be ensured even if the pressure fluctuates over a wide range from low pressure to high pressure, so the sealing performance can be maintained over a long period of time. It has become possible to provide a stationary double mechanical seal for high-pressure batch operation that is compact and economical and can be stably maintained.

[Brief explanation of drawings]

FIG. 1(a) is a schematic cross-sectional configuration diagram of main parts illustrating the mechanical seal according to the first invention, FIG. 1(b) is a schematic operational view of the same, and FIG. 2 is an example of the mechanical seal according to the second invention. FIG. 3(a) is a schematic cross-sectional configuration diagram of essential parts showing a conventional mechanical seal, and FIG. 3(b) is a schematic operational diagram thereof. 11.41... Retainer lla... Outer ring 11b.
... Inner ring (21a, 21b), (33c + 33d) ... Seat ring 31 ... Rotating shaft 32 ... Seal box 34a
+34b... Seat ring receiver 40a, 40b...
Driven Ring Applicant: Kobe Steel, Ltd. Nippon Virae Gyo Co., Ltd. Figure 1 (a) Figure 1 (b) 1 Figure 2

Claims (1)

[Scope of Claims] (11) Opposite side end surfaces of each seat ring which are loosely pin-coupled with an elastic body interposed to each seat ring receiver arranged at a predetermined interval in the seal box of the rotating shaft; , each sealed end surface is sealed by contacting the rear side end surfaces of each driven ring integrally disposed on the left and right driven ring holders of the ring-shaped retainer fixed to the rotating shaft. In the stationary double-acting mechanical seal, the ring-shaped retainer has an outer ring having both side surfaces formed into almost perfect planes, a thickness smaller than the thickness of the outer ring, and an outer diameter equal to the sheet. A stationary double-acting mechanical seal characterized by an inner ring having the same dimensions as the seal diameter of the ring receiver, which is slidably fitted in the axial direction. The opposite end faces of each seat ring are loosely pin-coupled with elastic bodies to seat ring receivers arranged in the same direction, and driven ring retainers on the left and right sides of the ring-shaped retainer fixed to the rotating shaft. In a stationary double-acting mechanical seal in which each sealed end surface is formed by contacting the rear side end surfaces of each driven ring that are integrally arranged, the ring-shaped retainer is almost completely sealed. An inner ring, which has a thickness smaller than that of the outer ring and whose outer diameter is the same as the seal diameter of the seat ring receiver, is slidably fitted in the axial direction to the outer ring, which has both sides formed in a flat plane. The seat rings are not inserted, and each of the seat rings is divided in the direction perpendicular to the axis and loosely connected with pins and abutted,
A stationary double-action mechanical seal characterized in that a sealing member is arranged on the corresponding contact surface so that the outer diameter of the seal is the same as the outer diameter of the sealed end surface.