JPH0320627B2 - - Google Patents

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 contact condition of the sealed end surfaces deteriorates, resulting in a reduction in sealing performance, which is a problem. That is, the configuration of a typical double seal of this type is shown in Figure 3a.
The explanation will be as follows based on (schematic vertical cross-sectional view of main parts). In the figure, 31 is a rotating shaft, and 32 is a seal box for the rotating shaft 31. Also 33a,
Reference numeral 33b denotes seat rings disposed within the seal box 3 at predetermined intervals. In this arrangement, the seat ring receivers 34a and 34b fixed at both ends inside the seal box 32 have a plurality of pins 36 and springs 3, respectively, along the circumferential direction.
7 are inserted alternately. Nao pin 3
6 is fitted into each pin hole to form a loose pin connection state. That is, seat ring 3
The entirety of 3a and 33b are seat ring receivers 34, respectively.
It is attached so that it can be moved back and forth and up and down to some extent with respect to a and 34b. The seat rings 33a and 33b are replaced by a driven ring 40a, which will be described later.
40b to form a double sealed end surface.

The driven rings 40a and 40b are both connected to the rotating shaft 3.
1 by integrally attaching the rotary shaft 31 to
The mounting structure is as follows. That is, both side surfaces of the retainer 41 are formed to be completely flat, and the opposing side end surfaces of the driven rings 40a and 40b are brought into contact with these flat surfaces via detent pins (not shown in the figure), respectively. The driven rings 40a and 40b are integrally disposed on the left and right sides of the retainer 41 by an upper driven ring holder 44 and a lower driven ring holder 43. In addition, both sides of the retainer 41 are lapped in advance to create an ultra-precise flat surface (in terms of surface roughness, μ=
It is finished to a flatness of about 0.0003 mm (about 0.7 μm in terms of flatness). In addition, driven rings 40a, 40b
Needless to say, the opposite end faces and the sealed end faces are lapped. In this way, the driven ring 40
The retainer 41 having left and right portions a and 40b is further fitted onto the stepped portion 31a of the rotary shaft 31 for positioning, and is fixed to the rotary shaft 31 by a sleeve 45 and a screw 46.

Further, 47a and 47'a are inlets for high-pressure cooling liquid (hereinafter referred to as sealing liquid), and 47b is an outlet for sealing liquid.
A sealing liquid is filled 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 prevent thermal distortion at the sealed end surface as much as possible, and the seat rings 33a and 33b are driven respectively. Consideration is given to sufficiently pressing the rings 40a, 40b 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 except 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 seals between the sealing liquid and atmospheric pressure. This is an O-ring that was arranged with this 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 due to the thrust force (indicated by the arrow in the figure) that acts on the retainer 41 when the pressure of the fluid to be sealed in the space 49 increases, as shown in FIG. This is due to the fact that the axial force (e.g., This is because the degree of distortion occurs with a similar distribution, and the gas-side sealed end face becomes a so-called open-face state as shown in the figure, resulting in a rapid increase in sealed fluid leakage.

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 ensure long-term stability of sealing performance without making the structure of the mechanical seal part particularly large, and we filed an application for utility model registration first. (Jitsugan No. 130766, Showa 58).
However, it has become clear that a new problem arises when equipment to which the dual seal of the earlier application is applied is operated in a so-called batch manner. In other words, from the start of operation to the specified pressurization operation (while the pressure is rising)
During the period from the predetermined pressurization operation to the stoppage of operation (while the pressure is decreasing), the end surface of the driven ring on the retainer side and the side surface of the sealed end surface of the retainer cannot properly face each other, so the sealed end surface is in a double-open state. As a result, it has become clear that leakage of sealing fluid still occurs, and particularly when the above-mentioned batch operation is performed at a high pressure, the range of pressure fluctuation is large, so that the problem of increasing the amount of leakage occurs. As a countermeasure to this problem, it may be possible to suppress the leakage of the sealed fluid within the permissible range by increasing the rigidity of the retainer and driven ring of the multiple seal according to the earlier invention of the present inventors, but this method is not possible. Since the shape becomes large and the material is expensive, 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 batch operation. be.

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

The configuration 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 basic configuration that is the same as the conventional example. .

FIG. 1a is a diagram corresponding to the main part explanatory diagram of FIG. 3b, and shows the set state of the multiple seal before pressurizing operation of the chemical reaction vessel. The retainer 11 is
It can be said that the structure is similar to that of a conventional retainer (41 in FIG. 3) divided into two in the circumferential direction. That is, the retainer 11 has an outer ring 11a fitted into an inner ring 11b so as to be slidable in the axial direction, and both sides of the outer ring 11a are almost completely flat (precision finished surfaces similar to conventional retainers).
is maintained, and the inner ring 11b has a thickness smaller than that of the outer ring 11a, and is configured such that its outer diameter is the same as the seal diameter Dφmm of the seat ring receiver 34b. As a result, the forces acting on the outer ring 11a of the retainer 11 are both (pressure P' of the sealing fluid + pressing force S of the spring 37), so the two acting forces are perfectly balanced, and the outer ring 11a
No moment is applied to either side of a. On the other hand, the pressure P of the sealing fluid acts on the sealing fluid side end surface of the inner ring 11b of the retainer 11, and its magnitude F is F=π/4(D ² - d ² )×P(Kgf/mm ² ) [ However, d is the inner diameter (mm) of the inner ring 11b, D is the outer diameter of the inner ring 11b, and the seat ring receiver 34
b is the seal diameter (mm)].

Due to this acting force, the inner ring 11b is distorted as shown in FIG. 1b, but this is because 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. In addition, both side surfaces of the inner ring 11b in the distorted state can be made to hardly contact the opposing end surfaces of the driven rings 40a, 40b.

Therefore, the outer ring 11 occupies the main part of the retainer 11.
Since a is not affected by the pressure P of the fluid to be sealed, the vertical flatness of the sealed end face is maintained sufficiently uniform even during pressure fluctuation operation under batch operation, and good sealing performance can be maintained.

However, even with such a double seal according to the first invention, when batch operation is performed under pressure conditions of about 100 kgf/cm ² or ^more , the sealing performance cannot necessarily be said to be good. 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. 1b, since the seat rings 33a and 33b are generally made of an ultra-hard material such as WC, their rigidity is lower than that of the driven rings 40a and 40b made of a soft material such as carbon. can be said to be extremely high. For this reason
A pressure of about 100Kgf/cm2 or ^less is applied to the driven ring 40.
Even if it acts on a, 40b, the retainer or driven ring 4
The problem of distortion generation as in 0a and 40b was almost negligible. Therefore, under such pressure conditions, the retainer 41 and the driven rings 40a, 40b
It should be possible to maintain good sealing performance as long as no distortion occurs, and as described above, the first invention was completed as a result of conducting studies based on this policy. However, the sealing fluid pressure P′
At 100 Kgf/cm ² or more, especially around 200 Kgf/cm ² , the moment in the direction of the arrow acting on the ends A and B of each seat ring 33a and 33b closer to the retainer cannot be ignored, and the moment at the ends A and B becomes more or less It was found that distortion occurred. As a countermeasure to this problem, it is possible to increase the size of the seat rings 33a and 33b, but this would result in an increase in the size of the seal structure and lead to an increase in manufacturing costs. Therefore, seat rings 33a and 33b
In order to prevent the above moment from occurring at the ends A and B closer to the retainer without making them particularly large, it is necessary to create a structure in which the ends A and B are not affected by pressure, that is, are balanced in terms of pressure. I was able to get the guidelines that I should do it.

The second invention was made as a result of further tests based on the above-mentioned guidelines, and it was found that even during pressure fluctuation operation of chemical equipment that is batch-operated at about 100 Kgf/cm ² or more, especially around 200 Kgf/cm ² It is an object of the present invention to provide a compact and economical dual seal that can maintain good sealing performance.

However, such a double seal of the second invention is as follows:
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 an outer ring that has both sides formed into almost perfect planes. The seat rings are inserted to form a retainer, and each seat ring is divided in the direction perpendicular to the axis, and they are loosely connected with pins to abut each other, and the corresponding contact surface is made so that the outer diameter of the seal is the same as the outer diameter of the sealed end surface. The gist lies in that a sealing member is arranged.

The configuration and effects of the second invention will be described below with reference to the drawings of the embodiments, and the second invention will also be described with a focus on its characteristic configuration.

FIG. 2 is a drawing corresponding to the cross-sectional view of the main part of FIG. 1a, and shows the set state of the multiple seal before pressurizing operation of the polymerization reaction vessel. seat ring 2
1a and 21b are both divided in the direction perpendicular to the axis, and have sliding parts 21a ₁ and 21b ₁ and supporting parts 21a ₂ and 21b _2.
is formed, and the structure of each pair of sliding parts and supporting parts is such that they sandwich the O-rings 23a, 23b (while being sealed by the O-rings) and abut at the lower part, and the pins 22a, 22b at the upper part. are loosely coupled via the O-rings 23a,
The clamping position of 23b is configured such that the seal diameter d ₀ φmm is the same as the outer diameter of the sealed end surface. As a result, each of the seat rings 21a ₁ and 21b ₁ receives pressure (sealing fluid pressure P') equally from both sides, so that no moment is generated in the direction of the arrow shown in FIG. 1b. Furthermore, each seat ring 21a ₁ ,
Even if 21b ₁ is slightly distorted in the radial direction, this does not pose a problem since it does not affect the vertical flatness of the sealed end face. In this way, retainer 1
In addition to the basic effect of the pressure balance method by circumferential division of 1 (the effect of maintaining the vertical flatness of the sealed end face by preventing distortion from occurring in the driven rings 40a, 40b), each seat ring 21a, 21b
The basic effect of the pressure balance method by dividing in the direction perpendicular to the axis (the effect of maintaining the vertical flatness of the sealed end face by preventing moment from being generated in the sliding portions 21a ₁ and 21b ₁ ) can be enjoyed. Therefore, as described above, good sealing performance can be maintained even during pressure fluctuation operation of the polymerization reaction vessel, which is operated in batches at about 100 Kgf/cm ² or more, particularly around 200 Kgf/cm ² .

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 be applied to the rotating shafts of all kinds of industrial machinery, automobiles, aircraft turbines, etc., and the sealing fluid can also be an aqueous solution. ,
Lubricating oil, gas, etc. can all be applied.

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 stabilized over a long period of time. It has now become possible to provide a compact and economical stationary double mechanical seal for high-pressure batch operation that can be maintained at high pressure.

[Brief explanation of the drawing]

FIG. 1a is a schematic cross-sectional configuration diagram of main parts illustrating the mechanical seal according to the first invention, FIG. 1b is a schematic operational diagram of the same, and FIG. FIG. 3a is a schematic sectional view of the main parts of a conventional mechanical seal, and FIG. 3b is a schematic operational view of the same. 11, 41...Retainer, 11a...Outer ring, 1
1b...Inner ring, 21a, 21b, 33c, 33d
... Seat ring, 31 ... Rotating shaft, 32 ... Seal box, 34a, 34b ... Seat ring receiver, 40a, 40b ... Driven ring.

Claims (1)

[Scope of Claims] 1. 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 a seal box of a rotating shaft; The stationary ring-shaped retainer fixed to the rotating shaft has its left and right driven ring retainers in contact with the rear side end surfaces of each driven ring integrally arranged, thereby forming each sealed end surface. In the dual-type mechanical seal, the ring-shaped retainer has an outer ring having both side surfaces formed into almost completely flat surfaces, and has a thickness smaller than that of the outer ring and an outer diameter that is the same as the seal diameter of the seat ring receiver. A stationary double-acting mechanical seal characterized in that an inner ring having dimensions of 2. The opposing end surfaces of each seat ring, which are loosely pin-coupled with an elastic body to each seat ring receiver arranged at a predetermined interval in the seal box of the rotating shaft, and the seat rings fixed to the rotating shaft. In a stationary double-acting mechanical seal, each sealed end surface is formed by contacting the end surfaces of each driven ring integrally arranged with the driven ring holders on the left and right sides of a ring-shaped retainer, thereby forming each sealed end surface. The ring-shaped retainer has an outer ring having both side surfaces formed into almost completely flat surfaces, and an inner ring having a thickness smaller than the thickness of the outer ring and an outer diameter having the same dimensions as the seal diameter of the seat ring receiver. The seat rings are fitted so as to be slidable in the axial direction, and furthermore, each of the seat rings is divided in the direction perpendicular to the axis and is loosely connected with a pin to abut, and the outer diameter of the seal is the same as the seal end surface of the corresponding contact surface. A stationary double-acting mechanical seal characterized by a sealing member arranged so as to have the same dimensions as the outer diameter.