RU65159U1 - HYDRODYNAMIC SEAL OF ROTATING COMPOUNDS (OPTIONS) AND SYSTEM OF HYDRODYNAMIC SEALS OF ROTATING COMPOUNDS - Google Patents

Abstract

A geometrical form for hydrodynamic lubrication of a dynamic sealing, mainly o-ring lip seal of a rotating joint, which is used to separate the lubricant and the environment. A dynamic sealing lip is necessary for establishing sealing contact with compression with a surface rotating relative to it and for driving a film of lubricating fluid into the joint between the dynamic edge and the rotating surface during relative rotation, which can be directed either clockwise or counterclockwise. A lift containing an elongated recess provides a gradual narrowing, an effective collision angle and a gradual increase in pressure in the contact zone, which contributes to effective hydrodynamic wedging. The tapered contact zones with increased pressure as a result of the edge effects of compression provide controlled movement of the lubricant inside the moving seal joint between the seal and the surface rotating relative to it, which improves lubrication and provides less torque required to overcome seal friction.

Description

This utility model relates generally to the lubrication of seals on rotating joints to extend the life of the system and reduce the torque needed to overcome seal friction, wear and the amount of heat generated. Rotary joint seals are suitable for isolating from the environment and retaining lubricant. In particular, the present utility model relates to the lubrication of seals, which can be subject to large pressure drops, large compression in percentage terms and can rotate both clockwise and counterclockwise.

In accordance with one preferred embodiment of the present utility model, the geometrical shape of the lip of a hydrodynamically lubricated seal is suitable both for compression seals of rotary joints (i.e., seals fitted with interference fit) and for flexible seals of rotary joints of seals with a lip projecting in the form of a cantilever.

Sealed downhole thrust bearings of downhole motors, rotary adjustable tools, rotary blowout preventers and other oilfield equipment are a variety of devices with difficult conditions for sealing rotary joints, as described in assigned US Pat. 6036192, 6109618, 6120036, 6227547, 6315302, 6334619, 6382634, 6494462 and 6561520. Examples of hermetic sealed thrust bearings of downhole motors are contained in assigned US Pat. 4372400, 4476944 and 6439866. Examples of rotating blowout preventers are described in assigned US patents 4281724, 5178215, 5224557, 5647444, 5662181, 6016880, 6354385 and 6554016.

The current state of the art in the field of seals for such equipment is represented by a family of hydrodynamically lubricated seals that are manufactured and delivered according to the aforementioned assigned patents. These seals, known in the industry under the trade name "KALSI SEALS", typically contain sine-wave lifts that are designed to introduce a film of lubricant into the moving seal joint between the dynamic seal surface and the rotating member. Though

such seals function well under certain conditions; there are other conditions in which their performance is not good enough. One set of such difficult conditions is when the seal operates in a clockwise or counterclockwise rotation and an increased film thickness of the lubricant is needed. An example application with such conditions is a downhole turbine engine. The seals of the rotating joints of the downhole motor operate under clockwise rotation during normal drilling and counterclockwise rotation under free rotation, which may occur when lowering the drill pipe string.

In existing seals containing lifts of a sinusoidal hydrodynamic shape, the film thickness takes on the greatest importance in the sinusoid region, since most of the lubricant that enters the front edge of the lift simply flows from the rear edge of the lift. As a result, the moving sealing joint is less well lubricated from the side facing the external environment, which leads to undesirable wear, torque and heat generation in the seal, especially if one or more of the conditions from the following list is present:

- lubricants with low viscosity, the value of which is less than the optimal value for lubrication;

- a large pressure drop across the seal;

- high geothermal ambient temperature;

- materials and shaft configurations that impede heat transfer;

- high initial compression of the seal;

- high rotation speeds;

- closely located additional seals, in which the heat generated in the seal is accumulated.

The torque during operation is associated with the shearing action of the lubricant and the contact of the roughness in the moving sealing joint. Although existing hydrodynamic seals operate with much less heat than non-hydrodynamic seals, the heat generated due to the presence of torque to overcome the friction of the seal continues to receive significant attention. Existing seals are usually made of elastomers, which are characterized by accelerated deterioration at elevated temperatures. For example, problems of resistance to the medium, problems of gas penetration, swelling, permanent compression deformation, extrusion

pressure damage - all problems are exacerbated at high temperatures.

Existing seals cannot be used in some installations with high speeds and high pressures, simply because the temperature is outside the usable temperature range of the sealant material due to the heat generated in the sealant. In addition, existing seals rotating in both directions do not provide sufficient lubrication if the sealing lip contains harder materials, such as reinforced plastic based on polytetrafluoroethylene (“PTFE”).

Typically, the life of the seal is ultimately limited by susceptibility to permanent compression (i.e., permanent deformation), abrasion, and extrusion damage. In many installations it would be useful to use hydrodynamic seals capable of handling large initial compressions, allowing the seal to withstand large mechanical deviations from the axis, runout, tolerances, permanent compression deformation and wear. In many installations, it would be preferable to have an improved lubricated seal under extremely harsh operating conditions and a seal that is subject to less wear. Also, in many installations it would be useful to use a seal that produces less heat, which would experience less loss of elasticity associated with temperature, and therefore would increase their extrusion resistance at high pressure.

10-13D from US Pat. The gaskets made in accordance with FIGS. 10-13D of US Pat. seal families for various shaft diameters.

The shaft diameter significantly affects the degree of narrowing for a given size of the convex hydrodynamic inlet section. Shafts with a smaller diameter taper with the inlet section at a higher speed, forming a sharper taper. This affects the effectiveness of clogging the lubricant. For the same size of the convex inlet portion, the seals with the smaller diameter shown in FIG. 10-

13D of US Pat. No. 6,109,618, are less lubricated and washed, and larger diameters are better lubricated and washed.

On Fig schematically shown in an uncompressed state, the relationship between the different diameters of the shafts and the radii of the oriented along the circumference of the hydrodynamic inlet sections. Figure 17 is not an image of the established narrowing, it gives an idea of the relationship between the diameter of the shaft and the narrowing for the input section.

As can be clearly seen from Fig. 17, the narrowing of the convex inlet portion of radius 2.00 "to the shaft with a diameter of 2.75" is approximately two times larger than it would be for a shaft with a diameter of 16.50 ". It is also shown that the narrowing of the convex inlet portion of radius 2.00" to a shaft with a diameter of 0.50 "is even larger. As a result, a shaft with a diameter of 16.50" is characterized by better lubrication and a higher washing intensity compared to shafts with a diameter of 0.50 "or 2.75" with the same size of the inlet section.

As can be seen from the dashed line in FIG. 17A, if the narrowing of the inlet portion with a radius of 2.00 "and a shaft with a diameter of 16.50" is doubled for a shaft with a diameter of 0.50 ", then it is understandable (as a result of recent studies cited by the authors of this utility model under contract No. DE-FG02-05ER84206 with the US Department of Energy) that the radius of the inlet should have been concave rather than convex, so small diameters should have concave inlets to achieve the same level of hydrodynamic lubrication as the larger Larger diameters with convex inlet portions. US Pat. No. 6,109,618 does not provide for concave inlet portions. In fact, methods for manufacturing tools for seals of FIGS. 10-13 of US Pat. No. 6,109,618 cannot be adapted for concave inlet portions.

Since the downhole equipment and tools are small, a number of issues need to be resolved, because some parameters cannot be reduced linearly due to practical production constraints, tolerances and specific requirements of the installations.

Since the seals of rotating joints are reduced, a decrease in the width of the cross-section along the radius leads to less compression in size at a given initial compression in percent. Tolerances cannot decrease below a certain real level, which contributes to the large boundaries of compression changes. The shaft stiffness also decreases non-linearly, contributing to greater radial deflection and runout. This results in more compression on one side of the seal and less compression on the other side.

A certain minimum level of compression is necessary so that without loss of sealing contact with the shaft, the seal can adapt to tolerances, deviation from the axis, wear of the seal, and permanent compression deformation. The factors formulated above prescribe that in order to achieve the required absolute compression value for miniature seals, it is necessary to increase the initial compression ratio expressed as a percentage compared to large diameter seals.

The pressure in the contact zone in the moving seal joint is associated with the compression, expressed as a percentage, and the elastic modulus of the seal material. A greater compression ratio leads to a higher pressure in the contact zone of the seal and the shaft, which reduces the thickness of the hydrodynamic lubricating film, and a decrease in the film in turn leads to higher friction, the release of more heat and more wear. The limits of compression for existing seals impose appropriate restrictions on the minimum cross-sectional dimensions practicable in production. To compensate for this, seals are required that can support a stronger hydrodynamic proppant in both directions of rotation.

Briefly, the present utility model relates to the shape of the generally annular hydrodynamically lubricated dynamic lip of the seal, which aims to increase the life of the seal in installations where the pressure of the lubricant can be significantly higher than the ambient pressure and where the environment may contain abrasive solids.

This utility model generally relates to seals of rotary joints designed to seal rotary components of a mechanism, while the seal must retain lubricant and prevent penetration from the environment, in addition, the seal interacts with the lubricant during relative rotation to clog the lubricant film material in a moving sealed joint between the seal and the surface rotating relative to it. More specifically, the present utility model relates to the particular hydrodynamic shape of the dynamic edge rises, which is suitable for rotating both clockwise and counterclockwise, and in order to achieve effective hydrodynamic lubrication, the present utility model controls the pressure in the moving sealing joint. This allows the dynamic lip to have a relatively high initial compression ratio and a relatively low torque to overcome seal friction and helps

withstand relatively high pressure drops.

Each of the lifts according to a preferred embodiment of the present invention contains an elongated recess that provides a gradual narrowing with a surface rotating relative to it in the circumferential direction, an effective collision angle and a gradual increase in the direction along the circumference of the pressure in the contact zone. These factors contribute to an effective hydrodynamic proppant. The beveled contact zones with increased pressure obtained due to the edge compression effects are needed for the controlled movement of the lubricant through the moving sealing joint in the direction of the environment, which is done for improved lubrication, for low torque to overcome seal friction during operation, for extended service life and for the ability of hydrodynamic functioning at low speeds. Some of the lubricant moves outside the contact areas with high pressure, thereby providing lubrication of the contact areas with high pressure. When the lubricant moves through a moving seal joint into the environment, the lubricant leaks.

The shape of the lifts according to the present invention is suitable both for compression seals (i.e., seals installed with interference) and for flexible seals of rotary joints with an edge projecting in the form of a cantilever. The mold may be configured for mechanical sealing when the dynamic lip protrudes substantially in the axial direction to contact a rotating surface of a planar shape. Alternatively, the mold may be configured for radial sealing when the dynamic edge protrudes substantially in the radial direction to contact a rotating surface of a cylindrical shape.

The shape of the inlet section, formed by the shape of the elongated recess, can be generally convex or concave, or in some places convex, and in other places, concave, which is determined by the diameter of the dynamic edge and the desired speed of narrowing with a rotating surface.

In order to understand in detail how the above-mentioned features, advantages, and objectives of the present utility model are achieved, the more detailed description that is briefly disclosed above is provided with reference to preferred embodiments of the utility model. These options are illustrated in the accompanying drawings. However, note here that the accompanying drawings illustrate only

typical embodiments of the present utility model and, thus, do not limit the scope of protection of the utility model, since the utility model allows other, equally effective, embodiments.

The drawings show:

figure 1 is a partial cross section showing the shape of the cross section of an annular hydrodynamic seal that implements the principles of the utility model;

figa is an enlarged partial shaded isometric view of the hydrodynamic seal of figure 1 in an uncompressed state and configured for radial compression, this view shows an elongated recess and two full lifting dynamic edges;

FIG. 1B is a partial plan view showing a lift of the seal of FIGS. 1 and 1A for clockwise rotation;

figs - section along CC in figv;

fig.1D, 1E and 1F - sections D-D, EE and F-F, respectively, in figv;

figure 2 is a view of the contact zone of two full rises of the dynamic edge of the seal of figa, this view contains a map of the contours of the relative pressure in the contact zone;

figa and 2B are views of alternative forms of the contact zone of the dynamic edge of the seal corresponding to the present utility model;

FIG. 2C is a view similar to that of FIG. 2, which shows a modified minimum deflection width and angle of contact zones with increased pressure, to improve lubrication and reduce leakage;

figure 3 is an enlarged partially darkened isometric view of a hydrodynamic seal, depicted in an uncompressed state and configured in the radial direction to seal a rotating surface having a cylindrical inner shape, such as a housing opening;

3A and 3B are partial isometric views in isometric of hydrodynamic seals depicted in an uncompressed state and configured axially to seal a rotating surface of a substantially flat shape, such as a shaft flange or housing;

4-9 are cross-sectional views of simplified embodiments of a hydrodynamic seal corresponding to the present utility model, these views are taken in the center of the elongated recess, while the seal is in uncompressed

condition;

10, 11 and 12-12B are cross-sectional views of preferred embodiments of a hydrodynamic seal corresponding to the present utility model, wherein the seals comprise spring-loaded spring elements for springing a dynamic edge relative to a rotating surface;

13 is a cross section of a preferred embodiment of a hydrodynamic seal, according to the present utility model, in the center of an elongated recess, wherein the seal is in an uncompressed state;

Fig. 14 is a partially darkened isometric view of the preferred embodiment of the hydrodynamic seal of Fig. 1, which is in an uncompressed state and configured for radial compression. This view depicts an elongated recess and two full elevations of the dynamic edge, the elongated recess has a greater depth to increase the temperature tolerance expansion and wear of the mold;

FIG. 15 is a partial isometric isometric view of a preferred embodiment of the hydrodynamic seal of FIG. 1, showing a counterclockwise lift for rotation and an elongated recess;

figa is an enlarged part of the hydrodynamic seal of Fig.15, showing the hydrodynamic inlet of the specified rise;

figv is a surface map of the hydrodynamic inlet section shown in figa;

Fig.16-16C is a schematic view illustrating a method of constructing map lines in an area limited by the construction profiles of Fig.15B, these map lines define the surface of the elongated recess shown in Fig.15 and 15A;

Fig.16D is a schematic view illustrating a construction method that can be used to set the narrowing speed of the variable curvature of the slope 49 of the recess and, thus, regulate the narrowing of the slope of the recess and the rotating surface;

17 and 17A is a schematic view illustrating, in an uncompressed state, the relationship between the shaft diameters and the radii of the hydrodynamic inlet section oriented along the circumference.

For a more complete understanding of the present utility model, FIGS. 1-2 B need to be considered together, since the general representation cannot be conveyed by any single figure. Elements in this description, which are denoted by the same reference position, have the same functions. For orientation, it should be noted that

for the cuts shown in figures 1, 1D-1F and 4-13, the cutting plane passes through the theoretical axis of the seal, the plane of the schematic views of Fig.16-17 is located at right angles to the theoretical axis of the seal, designed to seal the rotating shaft.

Figure 1 shows a partial cross-sectional view giving an overview of how the preferred embodiment of the utility model is used in the mechanism. The substantially annular seal 2 of the rotary joint comprises at least one dynamic lip 4, which also has an annular shape. The dynamic edge 4 corresponds to the special hydrodynamic features of the present invention (described together with the following figures), which allows the dynamic edge 4 to function with less heat generation, less wear, makes it more suitable for working on shafts with poor thermal conductivity and more suitable for use with strong the percentage of compression associated with small cross-sections of seals installed with interference.

The particular hydrodynamic features of the present utility model can be applied to the dynamic lips of many different types of seals of rotating joints, such as seals mounted with interference, or spring-loaded seals with an edge projecting in the form of a cantilever. To illustrate, the seal 2 shown in FIG. 1 is a typical seal installed with an interference fit that is installed in compression between the component 8 of the mechanism and the surface 5 rotating relative thereto.

The purpose of the seal 2 is to form a sealing contact with the rotating surface 5, to hold the volume of the first fluid 6, to separate the first fluid 6 and the second fluid 7, to isolate the second fluid 7 and to prevent the penetration of the second fluid 7 into the first fluid environment 6. In this description, the term "fluid" is used in a broad interpretation, covering both liquids and gases.

The first mechanism component 8 typically determines the location (i.e., location) of the seal 2. To illustrate a typical application, the first mechanism component 8 is depicted comprising a substantially annular seal groove 26 with a first groove wall 12 and a second groove wall 13 that are opposite each other friend. The conditions for the location of the seal may be different, which, however, does not go beyond the essence or scope of the utility model.

In the hydrodynamic seal industry, the first groove wall 12 is often referred to as the “lubricant side seal wall” and the second wall 13

grooves are often referred to as the “environmental seal wall”. Although the first groove wall 12 and the second groove wall 13 are shown to be stationary, with a constant relative position to each other, this does not limit the scope of the utility model, since other suitable shapes are allowed in the present utility model. For example, the first groove wall 12 and / or the second groove wall 13 may be separated from the component 8 of the mechanism for ease of maintenance and repair, and then installed in a more or less fixed manner so that a seal is located between them. In another example, if the seal is a seal with a housing and an edge projecting in the form of a cantilever, such as shown in FIGS. 12-12A, then only one wall is typically used to position the seal. Such seals are usually installed in a hole machined with a plating.

At least part of the dynamic edge 4 is compressed by the rotating surface 5 of the second mechanism component 10. During operation, the rotating surface 5 rotates relative to the dynamic edge 4 of the seal 2 and relative to the first component 8 of the mechanism. The utility model is applicable in such installations where either the first component 8 of the mechanism, or the second component 10 of the mechanism, or both components rotate.

The generally annular seal 2 has a theoretical center line (not shown), and the cross section of FIG. 1 is a cross section along the center line taken in a cutting plane that passes through the theoretical center line; that is, the theoretical center line is located on the cutting plane. The direction along the circumference of the relative rotation is normal (perpendicular) to the plane of the cross section, and the theoretical axial line of the seal 2, as a rule, coincides with the axis of relative rotation.

The seal 2 preferably contains a first end portion 14 of the seal, which is usually directed towards the first wall 12 of the groove and the first fluid 6. It is also preferable that the seal 2 comprises a second end portion 15 of the seal, which is usually directed to the second wall 13 of the groove and the second fluid 7 In the hydrodynamic seal industry, the first end portion 14 of the seal is often referred to as the “end portion on the lubricant side” and the second end portion 15 of the seal is often called the “end portion with side of the environment. " The first end portion 14 of the seal is preferably disposed generally opposite the second end portion 15. The first end portion 14 of the seal and the second end portion 15 of the seal may have other specific shapes without departing from the spirit or scope of the utility model.

Although it is shown that the first end portion 14 of the seal is positioned with a gap relative to the first wall 12 of the groove, the particular hydrodynamic shape of this utility model is also suitable for seals made to withstand simultaneous contact with both the first wall 12 of the groove and the second wall 13 of the groove in accordance with the principles of technical solutions described in assigned US patents 5873576, 6036192 and 6315302.

The rotating surface 5 may take on an outward and inward directed substantially cylindrical surface as desired, with the seal 2 being radially compressed between the first component 8 of the mechanism and the second component 10 of the mechanism. In this case, the axis of relative rotation is located essentially parallel to the rotating surface 5. When the seal has a radial configuration, the dynamic lip 4 is compressed essentially in the radial direction, and the peripheral wall 11 is made essentially cylindrical, and the first wall 12 of the groove , the second groove wall 13, the first seal end portion 14 and the second seal end portion 15 may be substantially flat.

In a preferred embodiment, the rotating surface 5 may be substantially flat, wherein the seal 2 is axially compressed between the first component 8 of the mechanism and the second component 10 of the mechanism. Such a solution is called “surface sealing”, in which case the axis of rotation is directed perpendicular to the rotating surface 5. With the axial (surface) shape of the seal, the dynamic lip 4 is compressed essentially in the axial direction, and the peripheral wall 11 can be essentially flat and the first groove wall 12, the second groove wall 13, the first seal end portion 14, and the second seal end portion 15 may be substantially cylindrical.

Compression (i.e., squeezing, contact) of the dynamic edge 4 by the rotating surface 5 establishes and defines the contact zone 16 at the interface between the dynamic edge 4 and the rotating surface 5. The contact zone 16 comprises a non-ring-shaped first edge 17 of the contact zone directed towards the first fluid 6, and the second edge 18 of the contact zone, preferably an annular shape, directed towards the second fluid 7.

The dynamic edge 4 preferably contains a dynamic stop line of intersection 19 of a steep annular shape, which is essentially aligned with the direction of relative rotation and is designed to prevent the penetration of the second fluid 7. The specified line of intersection 19 is made in

in accordance with the principles set forth in US Pat. No. 4,610,319. The dynamic stop line of intersection 19 has a shape that does not imply a hydrodynamic wedging action with the relative rotation of the dynamic edge 4 and the rotating surface 5. The dynamic stop line of intersection 19 is a stripping edge that helps to eliminate the penetration of contaminants from the contact zone 16 at the junction in case of any relative movement perpendicular to the relative direction rotation of the dynamic edge 4 and the rotating surface 5. A dynamic stopping line of intersection 19 is not necessary until, for more effective protection against penetration from the environment, a breakage and an annular shape are required.

The rotating surface 5 of the second component 10 of the mechanism and the peripheral wall 11 of the first component 8 of the mechanism are located at some distance from each other. For a compression seal, such as that shown in FIG. 1, the distance between the rotating surface 5 and the peripheral wall 11 is selected so that the seal located between them is compressed. Similar to conventional tightened seals, such as an O-ring or an O-ring self-aligning lip seal, compressing the seal 2 establishes a seal joint between the static sealing surface 20 of the seal 2 and the peripheral wall 11 of the first mechanism component 8. Compression also establishes the aforementioned contact zone 16 at the junction between the dynamic edge 4 and the surface 5 rotating relative to it.

When the dynamic edge 4 and the surface 5 rotating relative to it rotate relative to each other, the contact zone 16 is a moving sealed joint, while the dynamic edge 4 slides along the surface rotating relative to it 5. When the dynamic edge 4 and the surface 5 rotating relative to it do not rotate relative to each other, the contact zone 16 is a static sealed joint.

As shown, the peripheral wall 11 may be arranged substantially parallel to the rotating surface 5, or angled with the rotating surface 5, as shown in FIGS. 4, 6, 7, 8, or 9 of the assigned US Pat. No. 5,230,520.

In a preferred embodiment of the present invention, the width of zone 16 is different at different points of the circle in accordance with the particular hydrodynamic geometric shape shown in the following figures. The first edge of zone 17

The contact preferably has a non-annular (i.e. wavy) shape in accordance with a particular hydrodynamic geometric shape and, in response to the relative rotation of the seal 2 and the rotating surface 5, generates a hydrodynamic proppant. This hydrodynamic wedging action for lubrication drives the lubricating fluid film (i.e., the first fluid film 6) into the contact zone 16 at the interface between the dynamic edge 4 and the rotating surface 5, which reduces wear, torque and the amount of heat released. Therefore, with the relative rotation of the dynamic edge 4 and the rotating surface 5, the dynamic edge 4 slides or hydroplans over the surface of the lubricating fluid film. When the relative rotation is stopped, hydroplaning is also stopped, and a stationary seal is again established between the dynamic edge 4 and the rotating surface 5, which occurs by initial compression of the dynamic edge 4 relative to the rotating surface 5.

Hydroplaning minimizes or prevents the usual wear and tear during dry friction and high friction, typical of conventional non-hydrodynamic rubber and plastic seals, thereby increasing the life of the seal 2 and the life of the rotating surface 5. Thanks to hydroplanning, higher speeds, greater compression and greater differential are possible pressure even when part of the dynamic edge 4 is made of plastic, such as reinforced PTFE.

If a seal with an improved geometric shape corresponding to the present utility model is mounted on an unsecured pressure compensation piston, then the reduced torque required to overcome the seal friction is unlikely to cause the piston to circle in the inner diameter of the mating body. Also, reduced torque reduces the likelihood of unwanted slippage of the seal 2 relative to the component 10 of the mechanism compared to existing seals. This is a very important consideration for large diameter seals. Slippage can be different in different places of the circumference of a large seal, which will lead to the appearance of local stresses that contribute to tensile ruptures. For less torque seals, such breaks are less likely.

The second edge 18 (sometimes called the "environmental side") of the contact zone 16 preferably has a substantially annular shape and is aligned with the possible directions of rotation of the dynamic edge 4 and the rotating

relative to the surface 5. The second edge 18 of the contact zone is formed as a result of compression by the rotating surface 5 of the dynamic stop line of intersection 19. Since the second edge 18 of the contact zone has a substantially annular shape and is aligned with respect to possible directions of rotation, it does not generate a hydrodynamic wedging action in response to the relative rotation of the dynamic edge 4 and the rotating surface 5. According to the technical solution described in the assigned patent US 4610319, this contributes to the exclusion The penetration of the second fluid 7.

The dynamic edge 4 may be made of any suitable sealing material or combination of materials, including elastomeric or rubber-like materials, including, for example, elastomers reinforced with carbon fiber or fabric, plastics and various materials forming composite structures, such as, for example, a plastic casing, attached to an elastomeric structure (an example of such a structure is contained in FIG. 13)

It is preferable to make a dynamic edge 4 of an elastic sealing material with high characteristics of wear and resistance to extrusion. In the embodiment of FIG. 1, the dynamic edge is provided with a spring element 21 designed to press the dynamic edge 4 against the rotating surface 5 and to control the seal deformation and the force arising in the contact zone of the dynamic edge 4 and the surface 5 rotating relative to it 5. Also, the spring element 21 can press the static sealing surface 20 against the peripheral wall 11, as shown in figure 1. Without going beyond the boundaries of the volume or essence of the utility model, the spring element 21 can take any of the possible forms known in the art, including, for example, elastomeric rings and various forms of the spring. The spring element 21 can also be located in the annular groove 22 of any suitable shape.

Simplified embodiments are possible in which one or more elements are omitted. For example, there may be no spring element 21, and the seal 2 may contain one or more flexible edges or the seal 2 may be monolithic and made of one material, as shown in Fig.6 and 7, respectively.

As shown in FIG. 1, the spring element 21 may be an elastic material, the elastic modulus of which may be equal to or different from the elastic modulus of the dynamic edge 4. For example, the spring element 21 may be made of an elastomer with greater resistance to permanent deformation when

compression than the elastomer from which the dynamic edge 4 is made, however, both elastomeric materials can have essentially the same modulus of elasticity. In another example, the elastic modulus of the elastic member 21 may be less than the elastic modulus of the dynamic edge 4 in order to regulate the pressure in the contact zone to achieve optimal lubrication levels and low torque, as described in assigned patent US 5738358.

The pressure in the contact zone of the dynamic edge 4 and the rotating surface 5 associated with it is one of the important factors regulating the amount of heat released in the seal, since it affects the thickness of the hydrodynamic film, which, in turn, affects the shear rate of the lubricant film and the degree of roughness that may be present between the dynamic edge 4 and the rotating surface 5. Also according to the present utility model, using seals with a low initial pressure in the contact zone and can achieve a high washing level.

The spring element 21 may be connected to the rest of the seal 2 or molded as part of the seal 2, and thus a composite structure will be formed or the spring element 21 may simply be a separate part, which is mechanically attached to the rest of the seal 2 in a known manner. other suitable types of elastic elements are shown in this description in FIGS. 10-13. In figure 1, the seal 2 is a compression type seal, but replacing the spring element 21 with a mechanical spring, the main idea can be used for flexible seals with an edge protruding in the form of a console, as shown in figures 10-12 B. Simplifying a useful embodiment model, you can omit the spring element 21, so that the seal will be continuous and monolithic in cross section, as shown in Fig.7-9.

The first mechanism component 8 and the second mechanism component 10 together typically form at least a portion of the chamber 23 in which the first fluid 6 is located and from which lubricant is supplied. The first fluid 6, according to a utility model, is used to lubricate a moving seal joint between the seal 2 and the second component 10 of the mechanism during their relative rotation. Preferably, the first fluid 6 is a liquid lubricant, such as synthetic or mineral oil, although other fluids, including grease, water, and various process fluids, may also be used in some embodiments. The second fluid 7 may be any fluid that acts during operation on the seal 2, such as any liquid or

a gaseous medium, including, for example, a lubricating medium, process substances, flushing fluid, and so on.

The dynamic edge 4 of the installed seal includes a first edge side 24 directed to the first fluid 6 and a second edge side 25 directed to the second fluid 7. In the hydrodynamic seal industry, the first edge side 24 is often referred to as the “lubricated side” of the lip, the second side 25 The edge is often called the "environmental side" of the edge. The first side 24 of the edge is located at some distance from the dynamic stop line of intersection 19 and the second side 25 of the edge.

In figure 1, the seal 2 is located inside the groove 26 for sealing, which occurs when the pressure of the first fluid 6 exceeds the pressure of the second fluid 7. With this pressure ratio, the hydrostatic force arising from the pressure of the lubricant and acting between the rotating surface 5 and peripheral wall 11, presses the seal 2 against the second wall 13 of the groove. Since the shapes of the second end portion 15 of the seal and the mating second groove wall 13 are complementary, the seal 2 is well supported by the second groove wall 13 everywhere except for the gap 31 existing between the first mechanism component 8 and the second mechanism component 10. Preferably, the gap 31, often referred to as the “extrusion gap”, is small so that the seal can more easily close the gap and, by means of the rigidity of the seal material, resist deformation.

It is known in the art that with a large pressure drop, a small portion of the sealing lip tends to swell or extrude into the extrusion gap, and pulsating pressure and beating can bend the extruded material, resulting in damage caused by extrusion.

Tests of an improved geometric shape lip corresponding to this utility model have shown that it generates less heat in the seal compared to existing serial seals complying with US Pat. No. 5,230,520, and thanks to it, the seal more easily closes large extrusion gaps and withstands large pressure differences due to this that the dynamic edge 4 has a higher remaining modulus of elasticity, which is the result of less heat.

It was previously mentioned that the utility model is suitable for both radial and axial compression structures. In the case of a very large diameter and, therefore, very flexible seals, the dynamic lip 4 can simply

be made with a generally directed inward construction as shown in FIG. 1A (for the convenience of lithium under pressure) and used to seal a rotating surface 5 with a cylindrical surface directed outward (that is, such as a shaft surface).

Large diameter seals are not rigid with respect to torsion and, therefore, with respect to the cross section of large diameter seals, it is possible to rotate 90 ° so that the dynamic lip 4 becomes substantially axis-oriented, the lip intended to seal the rotating surface 5 is essentially , flat shape, or can be rotated 180 ° so that the dynamic edge 4 becomes an external, essentially radially oriented edge, designed to seal the rotating surface 5, representing an inwardly oriented cylindrical surface (i.e., an opening of the housing).

The relative torsional stiffness of small diameter seals is much higher and, therefore, for small seals, the dynamic lip 4 must be made with either a radial or axial sealing shape, depending on what is required for a particular embodiment.

As a result, the seal can be used as either a mechanical seal or a radial seal, configuring (twisting or manufacturing) the dynamic lip 4 so that it represents either the inner or outer diameter, or was the seal face, while retaining the advantages of this utility model related lubricated as described in this document.

FIG. 1A is an enlarged partially darkened isometric view showing two full elevations of the dynamic edge 4 of the same seal 2 as shown in FIG. 1, but here the seal is shown in an uncompressed state and configured (by twisting or manufacturing) for radial compression. In the following description, the abbreviation “CW” is used to mean “clockwise” rotation, and the abbreviation “CCW” is used to mean “counterclockwise” rotation.

The first end portion 14 of the seal, the spring member 21, the dynamic stop line of intersection 19, and the dynamic surface 51 of the seal are shown in FIG. 1A for orientation. The dynamic lip 4 can usually be located opposite the protruding static sealing lip 36, usually an annular shape. The protruding static sealing lip 36 preferably contains a static stopping line of intersection 38 breaking

an annular shape, which is essentially aligned with the direction of relative rotation and is designed to prevent the penetration of the second fluid 7.

The protruding static sealing lip 36 provides a certain degree of compression symmetry, which reduces the likelihood of twisting of the seal 2, which, if there is no lip 36, is possible when the seal 2 is installed, which corresponds to the description of the assigned patent US 5230520 and 6767016. In the embodiment shown in figa, the diameter of the protruding static sealing lip 36 is greater than the diameter of the dynamic lip 4, and the static sealing surface 20 is preferably an inclined surface such as shown in fig.1D-1F, and it provides additional compression of the seal next to the static locking line of intersection 38.

The protruding static sealing lip 36 may take other forms without leaving the boundaries of the volume and essence of the utility model; for example, if desired, any shapes of static edges shown in assigned US Pat. Nos. 5,230,520 and 6,767,016 can be used. FIGS. 4, 8 and 12B of this description provide several alternative forms of the protruding static sealing lip 36.

The embodiments illustrated here comprising a protruding static sealing lip 36 can also be simplified, if necessary, by eliminating a protruding static sealing lip 36 — for example, as shown in FIGS. 5, 9, 12 and 12A.

The portion of the edge shape shown in FIG. 1A is typically a single, usually looking, hydrodynamic lift of the type described in assigned US Pat. No. 5,230,520, however, this portion contains an elongated recess 40 in the middle. model, included a visual representation of exactly the following: a usually looking hydrodynamic lift, but containing in the middle an elongated recess 40, which from the functional point of view is divided t geometrical shape to lift 42CW to rotate clockwise and lift 42CCW to rotate counterclockwise, which are preferably mirror images of each other. The elongated recess 40 preferably defines a hydrodynamic inlet portion for both raising 42CW for clockwise rotation and raising 42CCW for counterclockwise rotation.

When the elongated recess 40 is compressed by the rotating surface 5 (FIG. 1), both ends thereof are gradually approaching the rotating surface 5 when viewed in

direction along the circumference. Thus, a hydrodynamic inlet zone is formed, the lumen of which gradually disappears at the contact point of the dynamic edge 4 and the rotating surface 5, which, in response to the relative rotation of the dynamic edge 4 and the rotating surface 5, contributes to the effective clogging of the film of the first fluid 6 into zone 16 contact between them.

The torque during operation of the seal depends on the total area of the contact zone (i.e., the area of the shear zone of the lubricant) and the degree of roughness of the contact of the dynamic edge 4 and the rotating surface 5. Compared to existing seals corresponding to assigned US Pat. Nos. 4,610,319 and 5,230,520, the elongated recess 40 removes the seal material from the compression zone eliminates unnecessary high pressure areas in the contact zone and reduces the total area of the contact zone and the total force acting in the contact zone of the dynamic fittin g 4 and the rotating surface 5 - it is not necessary to reduce the minimum width of the nip 16. This circumstance in itself reduces both the torque during operation and the torque during release compared to the existing prior art.

The operating torque also depends on the lubrication efficiency of the contact zone 16. The elongated recess 40 contributes to a significant improvement in the lubrication of the contact zone 16, which will be explained in detail below. Due to the improved lubrication, the seals can operate at high pressures in the contact zone, which are associated with a large percentage of the compression required for miniature seals and / or large pressure values in the contact zone caused by large pressure drops along the seal and / or large pressure values in the contact zone caused by the use of materials with a high elastic modulus, such as reinforced plastics, for sealing.

For rotation in two directions, the dynamic edge 4 comprises at least one lift 42CW for clockwise rotation and at least one lift 42CCW for rotation counterclockwise. The first side 24 of the dynamic edge 4 is preferably beveled with respect to the possible directions of relative rotation of the dynamic edge 4 and the rotating surface 5 and defines the profiles of the lift 42CW for clockwise rotation and the rise 42CCW for counterclockwise rotation. Both hoists 42CW and 42CCW for clockwise and counterclockwise rotation are adapted for hydrodynamic self-lubrication with a film of the first fluid in response to the relative rotation of the dynamic edge 4 and the rotating surface 5 both clockwise and

counterclock-wise.

With the relative clockwise rotation of the dynamic edge 4 and the rotating surface 5, the clockwise rotation rise 42CW is adapted to deflect the film of the first fluid 6 towards the second edge side 25 (and through, as a rule, the annular seal strip 44 shown in FIG. .2) and into the second fluid 7. Thus, with relative clockwise rotation, hydrodynamic lubrication of the contact zone 16 is provided (see FIGS. 1 and 2).

With a relative counterclockwise rotation of the dynamic edge 4 and the rotating surface 5, the counterclockwise rotation 42CCW is adapted to deflect the film of the first fluid 6 towards the second edge side 25 (and through, as a rule, the annular seal strip 44 shown in FIG. .2) and into the second fluid 7. Thus, with relative counterclockwise rotation, hydrodynamic lubrication of the contact zone 16 is ensured (see FIGS. 1 and 2).

As shown in FIG. 1A, the clockwise rotation rise 42CW and the counterclockwise rotation rise 42CCW are preferably mirror images of each other and, as shown, are preferably adjacent to each other to prevent lateral leakage from the contact zone 16 into the chamber 23. The use of at least one 42CW hoist for clockwise rotation and at least one 42CCW hoist for counterclockwise rotation makes the dynamic edge 4 bidirectional in the sense that an effective hydrodynamic drag yvanie and effective deflection of the first fluid film 6 to the second side edge 25 is achieved as in the relative rotation in the clockwise direction, and when the relative rotation counterclockwise.

The exact number of 42CW hoists available for clockwise rotation and 42CCW hoists for counterclockwise rotation depends on the diameter of the seal 2. Larger seals use more lifts — they usually repeat periodically, and smaller seals use fewer lifts. Extremely small diameter bidirectional seals can even use one 42CW lift for clockwise rotation and one 42CCW lift for counterclockwise rotation. The exact number of available 42CW clockwise rotations and 42CCW counterclockwise elevations also depends on the degree of lubrication required for relative rotation in each direction.

For example, if the pressure drop is large for clockwise rotation and small for counterclockwise rotation, it is advisable to use more 42CW rises for clockwise rotation and fewer 42CCW rises for counterclockwise rotation. In seals with different numbers of 42CW clockwise rotations and 42CCW counterclockwise risers, there are places where the clockwise rotation rise is adjacent to another clockwise rotation rise or where the counterclockwise rotation rise is adjacent to another lift to rotate counterclockwise. Then, at the junctions, both lifts have the same cross-sectional profile, so that the transition from one lift to another is smooth.

The closest to the first side 24 of the edge of the parts of the hoist 42CW for clockwise rotation and the hoist 42CCW for counterclockwise rotation are the multifunctional edge 46CW and the multifunctional edge 46CCW, respectively. The shape of the multifunctional edges 46CW and 46CCW is more easily understood in FIGS. 1C-1F. Each multifunctional edge 46CW and 46CCW contains at least a beveled portion (i.e., not parallel) with respect to the direction of possible relative rotation. Preferably, almost all of the multifunctional edges 46CW and 46CCW are beveled with respect to the direction of possible relative rotation and with respect to the dynamic stop line of intersection 19. It is preferable that the beveled parts of the multifunctional edge 46CW are beveled in the opposite direction with respect to the beveled parts of the multifunctional edge 46CCW .

At least a portion of the multifunctional edge 46CW, 46CCW forms an inclined portion 48CW, 48CCW, respectively. Each inclined portion 48CW and 48CCW helps to form one side of the elongated recess 40. When the seal 2 is compressed by its rotating surface 5 (FIG. 1), it is preferable that at least a portion of each inclined portion 48CW and 48CCW is compressed by the rotating surface 5 and, at least a portion of each inclined portion did not contact the rotating surface 5, but was located with some clearance relative to the rotating surface 5. The inclined sections 48CW and 48CCW are smoothly inclined / curved with respect to the rotating eysya surface 5, and thereby a shape promotes formation of a hydrodynamic wedge.

It is also preferable that the elongated recess 40 contains a slope 49 of the recess having a cross-sectional profile of a convex curve (as the most

clearly visible in fig.1D). The radius of the profile of the convex curve is preferably different in different places along the length of the lift, as will be described in more detail when considering figv-1F. Also, the dynamic edge 4 preferably comprises roundings 50CW and 50CCW of the recess, which usually have a concave curve profile in cross section (as is best seen in FIG. 1D), and in the direction along the circumference has a beveled, inclined profile. Preferably, at least a portion of the 50CW, 50CCW rounding of the recess is recessed compared to the inclined portions 48CW, 48CCW, respectively. Preferably, the 50CW, 50CCW curves smoothly mate or essentially smoothly mate with the slope 49 of the recess. It is also preferable that the 50CW, 50CCW curves smoothly mate or substantially smoothly mate with the inclined portions 48CW and 48CCW, respectively. Curved profiles of the slope 49 of the recess, the curves 50CW, 50CCW of the recess and the inclined portion 48CW, 48CCW can take any suitable shape, such as part of a circle, part of an elliptic curve or part of a parabolic curve, a combination of curves, and so on.

During operation, the lubricant (first fluid 6) collects contaminants inside the mechanism. With relative rotation, a portion of the first fluid 6 and the contaminants floating therein rotate in a circle inside the chamber 23 and rotate relative to the dynamic edge 4. Due to the beveled shape of the multifunctional edges 46CW and 46CCW, a portion of the first fluid slides along the dynamic edge 4 rather than into the moving edge zone 16 contact. This deflecting action helps to direct the contaminants inside the first fluid 6 away from the elongated recess 40, which helps to prevent wear of the dynamic surfaces of the dynamic edge 4.

FIG. 1B is a partial top view showing the rise 42CW for clockwise rotation of the seal 2 shown in FIGS. 1 and 1A. Figure 1C is a section CC of Figure 1B. Figures 1D, 1E and 1F are sections D-D, EE and F-F, respectively, figv. 1D-1F, for orientation, the first end portion 14, the second end portion 15, the static sealing surface 20, the spring element 21, the annular groove 22, the first edge side 24, the protruding static sealing edge 36 and the static stop line of intersection 38 are marked. For In orientation, some of these elements are also indicated in FIGS. 1B and 1C. In FIG. 1B, the length of the primary arc 80, the length of the secondary arc 82, and the depth 94 of the primary arc correspond to the same reference numerals that will be described in detail in FIGS. 16 and 16A.

1B, the first end portion 14 is parallel to the viewing plane (i.e., the plane of the triangular projection). Thus, the inclined portion 48CW located at an angle to the first end portion 14 (as can be seen in FIGS. 1A and 1C) is shown in perspective in FIG. 1B. In this perspective view, it is clear that the inclined portion 48CW is inclined with respect to the dynamic seal surface 51. As shown in FIG. 1B, it is preferable that this oblique shape of the inclined portion 48CW is a complex curve consisting of a primary curve 52 and a secondary curve 54 that are smoothly mated at the interface T55. One example of the construction of the above complex curve is shown schematically in FIGS. 16 and 16A; this method automatically adjusts the curvature of the primary curve 52 to maintain the same constriction speed with the shaft for a family of seals of different diameters, the sizes of which are selected for different external dimensions of the equipment. The 48CW ramp is part of the 46CW multifunctional edge. The multifunctional edge 46CW preferably transitions from the cut D-D to the cut F-F as shown in FIG. In any case, irrespective of the shape of the multifunctional edge 46CW, the clockwise rotation 42CW and the counterclockwise rotation 42CCW are adjacent to each other in the common place of the elongated recess 40, unless there are different numbers of clockwise rotations; arrow and rises for counterclockwise rotation.

The primary curve 52, which is preferably smoothly mated with the dynamic seal surface 51, controls the narrowing of the inclined portion 48CW (and to some degree of curvature 50CW of the recess) and the dynamic edge 4 and the rotating surface 5 when viewed in a circumferential direction. The degree of narrowing is an important aspect of hydrodynamic performance, as it controls the increase in pressure in the circumferential direction in the contact zone (known as DPDX) of the dynamic edge 4 and the rotating surface 5.

Since the 50CW curvature of the elongated recess 40 is smoothly associated with the inclined portion 48CW and the slope 49 of the recess, the curvature of the inclined portion 48CW is one of the three main factors regulating the established narrowing of the elongated recess 40 and the rotating surface 5. The other two of the three factors are speed and the magnitude of the change in the variable curvature of the curve 50CW and 50CCW of the recess and the rate of change of the variable curvature of the slope 49 of the recess. These factors will be described in detail below.

The circumference of the 42CW hoist must change in the family

seals of different diameters, since the same length is not completely divided by each value of the circumference of the seal. The indicated change in the length of the lift was also encountered in technology in seals with sinusoidal lifts (it becomes stronger in seals of small diameter). In existing seals, a change in lift length affects the rate of narrowing of the seal and the rotating surface. Additional information on changing the narrowing of existing seals, which is a function of the length of the lift, is contained in US Pat. No. 6,382,634. A fundamental improvement in the prior art is that in a utility model, the narrowing remains constant over a range of seal diameters. This property will be described in detail with reference to FIGS.

The primary curve 52 can take any suitable form, which provides the formation of a hydrodynamic wedge, without going beyond the scope of the utility model. The shape can be part of a circle (including part of a circle of infinite radius, that is, a straight line), part of an ellipse, part of a sinusoidal curve, part of a parabolic curve, part of a cycloid, part of a versper / version, part of a combination of curves, part of a combination of curves and straight lines, and so on. Further.

The projection shown in FIG. 1C and the cross-sections in FIG. 1D, 1E and 1F show a preferred surface shape of the dynamic edge 4 corresponding to the present utility model. The dashed lines used in FIG. 1C show the location of the main points T56, T57, T58 and T59 of the interface between various parts of the surface of the dynamic edge 4.

When observing perpendicular to the cross section, that is, as shown in FIGS. 1D, 1E and 1F, the surface between the dynamic stop line of intersection 19 and T56 is preferably a substantially straight line, which, when uncompressed, can be either inclined as desired , or parallel to a given rotating surface 5.

When viewed perpendicular to the cross section, that is, as shown in FIGS. 1D, 1E and 1F, the surface of the slope 49 of the recess between T56 and T57 is preferably a convex curve. At least part of the distance between the DD line (FIG. 1D) and the EE line (FIG. 1E), the radius of the curve between T56 and T57 preferably increases non-linearly, and in some places the curve becomes a curve with an infinite radius (i.e., a straight line) line) or becomes a curve, which in practice differs from a straight line by small distances arising in ordinary

rotary joint seals. The indicated change in surface curvature between T56 and T57 is a preferred method of forming a slope 49 of the recess, although for simplicity, the curvature may be constant rather than variable.

For example, the curve between T56 and T57 may have a radius that increases from 3.81 mm to 10160 mm between the lines D-D and E-E, and between the lines E-E and F-F the radius of curvature can remain constant and equal to 10160 mm. Obviously, a radius of 10160 mm implies in practice that the curve is a straight line between T56 and T57, which is common for the width of a typical dynamic seal lip of a rotating joint.

The change in the radius of curvature between T56 and T57 ensures that the input side edges 60CW and 60CCW of the contact zone 16, which are shown in FIG. 2, are preferably made in a curved shape. In the example shown in Figs. a straight line and remains the same from line EE to line FF.

When viewed perpendicular to the cross section, that is, as shown in FIGS. 1D, 1E and 1F, the surface between T57 and T58 (rounding 50CW of the recess) is preferably concave. At least in part of the distance between the DD line (FIG. 1D) and the EE line (FIG. 1E), the radius of the usually concave curve between T57 and T58 preferably increases non-linearly, and to the EE line this curve becomes a curve with an infinite radius (i.e., a straight line) or becomes a curve that, in practice, approaches a straight line for distances common for sizes of a typical dynamic edge 4. In the example shown in FIGS. 1-1F, the radius of the usually concave curve between T57 and T58 increases (then there is curvature becoming smaller and smaller) from the DD line on line E-E, thus at line E-E curve approximately coincides with a straight line, and remains the same from the line E-E to the line F-F. To simplify, the radius of curvature of the 50CW curve can be constant, not variable.

Preferably, the radius of the substantially concave curve between T57 and T58 is completely divided by the radius of the adjacent, generally convex, curve located between T56 and T57 and representing the slope 49 of the recess. Particularly suitable is a portion of about 1/10 to 1/2, so that at any point in the cross section between the DD line (FIG. 1D) and the FF line (FIG. 1F), the curve between T57 and T58 is from about 1/10 to 1/2 of the radius of the curve between T56 and T57.

The dynamic seal surface 51, at least for the seal of Figs. which, from a practical point of view, approach straight lines, that is, they are straight lines for the distances usual for the dimensions of ordinary dynamic edges 4.

When observing perpendicular to the cross section, that is, as shown in FIGS. 1D, 1E and 1F, the profile of the multifunctional edge 46CW between the mating points T58 and T59 is preferably a substantially convex curve. One of the goals of the curved profile is to help achieve a gradual narrowing of the dynamic edge 4 and the rotating surface 5, in order to serve as a hydrodynamic inlet portion to provide a hydrodynamic proppant with a relative counterclockwise rotation of the dynamic edge 4 and the rotating surface 5. Another goal is to so as to create a zone 62CW contact with high pressure inside the zone 16 of the contact (see figure 2). The specified zone 62CW is designed to deflect the lubricant in the direction of the dynamic stop line of intersection 19 in the direction opposite to the direction of rotation. The dimensions of the multifunctional edge 46CW are selected so as to provide the necessary level of contact pressure in the contact zone 62CW with increased pressure inside the contact zone 16. If desired, the profile radius of the multifunctional edge 46CW can be changed along its length in order to regulate the contact pressure in the contact zone 62CW with high pressure inside the contact zone 16. The specified pressure affects the torque, the hydrodynamic suction effect in one direction of relative rotation and affects the deviation of the lubricant film in the opposite direction of relative rotation. For example, if it is necessary to reduce the size of the contact zone 62CW with the increased pressure of zone 16 to deflect a smaller amount of the film of the first fluid 6 towards the dynamic stop line of intersection 19 and the corresponding second edge 18 of the contact zone and to form a smaller hydrodynamic leak, it is possible to partially or completely the distance from the line EE (FIG. 1E) to the line FF (FIG. 1F) increase the radius of the generally convex curve between T58 and T59. In the example shown in FIGS. 1-1F, the radius of the generally convex curve between T58 and T59 is increased (that is, the curve has less and less curvature) between the line EE and the line F-F; on figs should be drawn

attention to the trend of discrepancy between T58 and T59, resulting from changes in curvature.

The above-mentioned curves between T56 and T57, between T57 and T58 and between T58 and T59 can take any suitable curve shape or part of a curve, or a combination of curves, or a combination of curves and straight lines (i.e. a curve with an infinite radius of curvature). For example, curves can be selected from a group of curves that contains circles, ellipses, sinusoids, parabolic curves, cycloids, and verpers / versions. However, in the example shown in FIGS. 1-1F, the curves between T56 and T57, between T57 and T58 and between T58 and T59 are parts of circles.

FIG. 2 shows in a plane a contact zone 16 for two full rises of the dynamic edge of the seal 2 of FIGS. 1-1F when the seal is mounted on the rotating surface 5. The clockwise and counterclockwise rotations indicated in FIG. 2 are oriented so , as if the viewer was looking at the contact zone, leaving a mark on the rotating surface 5 (and would not look at the dynamic edge 4).

As mentioned above in the description of FIG. 1, the contact zone 16 comprises a first edge 17 of the contact zone and a second edge 28 of the contact zone. When the dynamic stop line of intersection 19 of the dynamic edge 4 is compressed by the rotating surface 5, in general, the annular second edge 18 of the contact zone is substantially aligned with the possible directions of rotation and therefore, in response to the rotation of the second edge 18 of the contact zone is not the cause of the hydrodynamic proppant the action of the second fluid and, therefore, prevents the penetration of the second fluid 7 into the contact zone 16, as described in the description of the assigned patent US 4610319.

Figure 2 shows a map of the contours of the relative values of the contact pressure in the zone 16 of the contact. The contact pressure shown as outlines in FIG. 2 was established by the finite element method, simulating an axial load of 15 pounds per square inch through a washer (to prevent mowing, as described in Society document No. 37627) oil engineers and the International Association of Drilling Contractors). The highest contact pressure is observed at the 64CW and 64CCW places of the highest pressure, and the lowest contact pressure is observed along the first edge 17 of the contact zone. The shape of the contact zone and the contours with the same value of the contact pressure remain almost constant, such as shown in figure 2, at different simulated temperatures that are within the operating temperature range of conventional seal materials.

The pressure contact zones 62CW and 62CCW, which are beveled in opposite directions with respect to the possible directions of relative rotation, are formed as a result of compression of the multifunctional edges 46CW and 46CCW of the dynamic edge 4 by the rotating surface 5. The contact pressure is increased in the pressure contact zone 62CW and 62CCW due to beyond the classical edge compression effects, the cross-sectional shape of the multifunctional edges 46CW and 46CCW and the inevitable compression of some parts of the seal material located between many functional- edges 46CW and 46CCW and the first end portion 14 of the seal.

The pressurized contact zones 62CW and 62CCW are configured so that in the circumferential direction, the contact pressure level therein increases gradually to allow at least a certain amount of lubricant to pass through them both during clockwise rotation and during rotation counterclock-wise.

The first edge 17 of the contact zone contains the input side edges 60CW and 60CCW, which are formed as a result of the compression of the recess slope 49 by the rotating surface 5. At least a part of the input side edges 60CW and 60CCW, and preferably all of them, are beveled with respect to possible directions of rotation, except for the T73 mate point. The beveling causes one or the other inlet side edge 60CW and 60CCW to function as a hydrodynamic inlet portion in any direction of rotation. The compression of the groove slope 49 by the rotating surface 5 gives rise to contact pressure, which gradually increases in the circumferential direction, which can be seen due to the contact pressure contours in FIG. 2.

The first edge 17 of the contact zone preferably comprises circumferentially projecting edges 66CW and 66CCW that are circumferentially located after / outside the edges 68CW and 68CCW of re-entering the contact zone, respectively. The circumferentially protruding edges 66CW and 66CCW are formed by compression by the rotating surface 5 of the inclined portions 48CW and 48CCW, respectively. The re-entry edges 68CW and 68CCW are formed by compression by the rotating surface 5 of the recess slope 49 and the rounding 50CW and 50CCW of the recess, respectively.

Basically, the contact zone 16 can be divided into a contact zone 70CW for clockwise rotation and a contact zone 70CCW for counterclockwise rotation and a substantially annular seal strip 44, which is limited

the dashed boundary line 72 touching the first edge 17 of the contact zone at the point T73 mates. The clockwise rotation contact zone 70CW corresponds to a rise 42CW of the dynamic edge 4 and is formed by compression by the rotating surface 5 of the portion of the clockwise rotation rise 42CW. Similarly, the contact zone 70CCW for counterclockwise rotation corresponds to the rise 42CCW of the dynamic edge 4 and is formed by compression by the rotating surface 5 of the part of the 42CCW hoist for counterclockwise rotation. In general, the annular seal strip 44 is basically an annular region of the contact zone 16, which is located between the dashed boundary line 72 and the second edge of the contact zone 18 (i.e., in general, the annular seal strip 44 is located between the second edge of the contact zone 18 and the zones 70CW for clockwise rotation and 70CCW for counterclockwise rotation).

The contact zone 16 is characterized by a minimum entrance width 84 and a minimum deviation width 85, which may be equal or different in size. Preferably, the minimum width 84 of the entrance was less than the minimum width 85 of the deviation, the reasons for this will be described later when considering figs.

The rotating surface 5 is wetted by the first fluid 6, which is preferably a liquid oil lubricant. When rotated counterclockwise, the rotation speed captures a thin film of the first fluid 6 after the input side edge 60CW, the re-entry edge 68CW of the contact zone and the circumferential edge 66CW, and the contact zone 70CW for clockwise rotation, which is part of the contact zone 16. This introduction of the lubricant film results in a hydrodynamic lifting force in the contact zone 16, which leads to a slight increase in the dynamic edge 4, which eliminates most of the direct sliding friction between the dynamic edge 4 and the rotating surface 5. When the film is in the rotation contact region 70CW clockwise, the rotation speed forces the film to move mainly along the circumference in a clockwise direction inside the contact zone 70CW for clockwise rotation, and some part of the film and leaves the contact zone 70CW after the pressure contact zone 62CW and re-enters the chamber 23, which is shown in FIG. 1. The pressurized contact zone 62CW, which is beveled relative to the direction of relative rotation, slightly reduces the thickness of the film and causes the part of the film that moved around the circumference in the clockwise rotation zone 70CW to move into the annular seal strip 44 and towards the annular second

edge 18 of the contact zone. When the film enters the annular seal strip 44, it is replaced by another part of the lubricant entering the annular seal strip 44, and moves more or less in a spiral, leaving the contact zone 16, passing through the second edge 18 of the contact zone. The result is a leak of the first fluid 6 into the second fluid 7 and lubrication of the annular seal strip 44. Said movement of the first fluid 6 into the second fluid 7 provides a flushing action that helps to remove contaminant particles that may be present in the contact zone 16, and also helps prevent contaminants from moving towards the chamber 23. The amount of leakage and flushing action may vary from small to large values by changing the shape of the seal. So, changing the curvature of the primary curve 52, you can change the narrowing of the dynamic edge 4 and the rotating surface 5 in the elongated recess 40.

During rotation, the deflecting action lubricates the mating surfaces of the seal 2 and the rotating surface 5 over the entire contact zone 16 even under conditions where the pressure of the first fluid 6 can exceed the pressure of the second fluid 7 by hundreds or even thousands of pounds per square inch.

Upon clockwise rotation, after the pressure contact zone 62CW, the thin film of the first fluid 6 is also entrained in the contact zone 70CCW for counterclockwise rotation. When the film is in the anti-clockwise rotation zone 70CCW, the rotation speed forces it to move mainly in a clockwise circle until it leaves the anti-clockwise rotation zone 70CCW through the input side edge of 60CCW, the contact zone reentry edge 68CCW and the circumferential edge 66CCW. This substantially circular movement of the film in the counter-clockwise rotation zone 70CCW provides excellent lubrication of the counter-clockwise rotation zone 70CCW without a noticeable increase in torque to overcome the friction forces of the seal.

In accordance with the law of physics, two different objects cannot occupy the same space at the same time. Over the entire length along the circumference of the counterclockwise rotation zone 70CCW, the presence of a film moving generally around the circumference in said zone minimizes or prevents any lateral leakage from the substantially annular seal strip 44 to the counterclockwise rotation zone 70CCW and thus minimizes or prevents any lateral leakage into the first fluid 6 located at

chamber 23 and therefore helps to maintain sufficient film thickness inside the annular seal strip 44, which is necessary to ensure sufficient lubrication.

When rotating clockwise by substantially circularly moving the film inside the counterclockwise rotation contact zone 70CCW, a constant flushing action takes place in the counterclockwise rotation contact zone 70CCW that transfers the heat generated in the seal from the contact zone 16 into the main volume of the first fluid 6 located in the chamber 23. This helps maintain the temperature of the dynamic edge 4 close to the temperature of the first fluid 6.

In seals fitted with an interference fit, such as that shown in FIGS. 1-1F, the pressure contact zones 62CW and 62CCW, which are closest to the second edge 18 of the contact zone, are also in a region that is in line with the force compression caused by the contact of the static sealing surface 20 and the peripheral wall 11. Therefore, there is a tendency to increase the magnitude of the contact pressure in those parts of the contact zones 62CW and 62CCW of contact with high pressure, which are located closest to the second edge 18 of the zone of the contact ta. The preferred variable radius of curvature of the multifunctional edges 46CW and 46CCW shown in FIG. 1A helps to control the magnitude of the contact pressure in the portions of the pressurized contact zones 62CW and 62CCW that are located closest to the second edge 18 of the contact zone. To simplify, the curvature of the 46CW and 46CCW multifunctional edges can be fixed rather than variable. FIG. 1B shows the seal of FIG. 1A, wherein the curvature of the multifunctional edges 46CW and 46CCW changes in a sinusoidal fashion from a minimum value on line EE to a maximum on line F-F. The curvature of the multifunctional edges 46CW and 46CCW also changes in FIG. 3B and, as a simplification, does not change in FIGS. 3, 3A, 14 and 15.

As for the first fluid 6, which, when turned clockwise, enters the contact zone 70CW for clockwise rotation, the lower the pressure in the increased pressure contact zone 62CW, the less the first fluid will deviate to the second edge 18 of the contact zone and the lower the rate of hydrodynamic leakage of the first fluid 6 into the second fluid 7. However, with regard to the first fluid 6, which enters the contact zone 70CCW for counterclockwise rotation, the lower the pressure in the contact zone 62CCW higher pressure, the greater the amount of the first fluid 6 will fall into the zone

70CCW contact for counterclockwise rotation.

In general, from the point of view of reducing the torque required to overcome the seal friction, it is preferable that the hydrodynamic absorption property of the elongated recess 40 is as strong as possible and, therefore, as much lubricant as possible will enter the contact zone 16, and it is desirable that the 62CW zones and the 62CCW high pressure contacts were as active as needed to deflect the amount of film through the O-ring strip 44 of a seal sufficient for proper lubrication when provided ennyh operating conditions. The variable radius of curvature of the multifunctional edges 46CW and 46CCW, which are shown in FIGS. 1A, 1B and 3B, helps to adjust the deflection of the film by removing the seal material from the compression zone to control the pressure in the pressure contact zones 62CW and 62CCW.

The counterclockwise rotation contact zone 70CCW is preferably a mirror image of the clockwise rotation contact zone 70CW, which is performed to provide the same functions as the contact zone 70CW for the clockwise rotation when rotating in the opposite direction.

When rotating counterclockwise, a similar hydrodynamic lubrication occurs, as described above, given that the rotation is counterclockwise. In other words, the functions of the clockwise rotation zone 70CW and the counterclockwise rotation zone 70CCW change places when the direction of rotation is reversed.

In the seals described in US Pat. effects ”compression. When contact pressure is plotted, this high pressure looks like a wavy crest that runs along the lubricated edge of the contact zone.

In the present invention, a lubricant that is hydrodynamically driven by the wedge into the contact zone 16 in the elongated recess 40 does not have to intersect the ridge-shaped high-pressure site in the contact zone, since the elongated recess 40 and especially its curves 50CW and 50CCW of the recess can simply be shaped so that the narrowing with the rotating surface 5 was smooth enough, which provides control of the magnitude and profile of the contact pressure, while

formation of a contact pressure ridge is prevented and thus a high efficiency of a hydrodynamic proppant is achieved.

The low torque to overcome seal friction provided by the present invention results in less heat being generated, which helps minimize the decrease in elastic modulus of the seal material associated with temperature. This helps the seal to better resist extrusion and be less prone to the classic “breaking off small pieces” or “extrusion damage”, as confirmed by comparative testing. Since extrusion damage can contribute to the absorption of the abrasive and can also disrupt the film of lubricant near the annular stopper shape and lead to additional friction and wear, the reduction in torque provided by the present invention greatly contributes to a longer service life under adverse operating conditions.

In some equipment (such as sealed downhole motor bearing devices), relatively high shaft deflection and other factors allow significant radial movement between the shaft and the seal housing, resulting in increased extrusion damage. The increased extrusion resistance corresponding to this utility model will help extend the life of the seal of such equipment. The improved lubrication provided by the utility model also helps withstand the high level of initial compression that is required to adapt to high levels of radial movement between the shaft and the seal housing.

At a low pressure drop, laboratory tests without cooling with a lubricant with a viscosity coefficient of ISO 32 showed that for seals made in accordance with this utility model, the torque is halved compared to existing seals made in accordance with figure 2 of US patent 5230520 Also, seals made in accordance with this utility model are able to rotate twice as fast as existing seals before a critical temperature is reached. temperature associated with the release of heat in the seal. The seals were tested on separate parts of the same shaft, with the same runout and the same seal body, while the seals were not cooled by anything but stationary air at room temperature.

The design technique for the geometric shape of the dynamic edge 4 described above when considering FIGS. 1-1F is extremely flexible and can be easily

modified to obtain the necessary modifications to the shape of the contact zone and the pressure profile in the contact zone.

For example, the shape of the contact zone 16 shown in FIG. 2A can be obtained by increasing the length along the circumference of the elongated recess 40. In another example, the shape of the contact zone 16 shown in FIG. 2B can be obtained by reducing the curvature of the primary curve 52 so that the primary curve 52 took along the circumference most of the 42CW hoist for clockwise rotation or the 42CCW hoist for counterclockwise rotation.

For maximum lubrication, the shape of the contact zone 16 shown in FIG. 2 is more preferable than the shape of the contact zone 16 shown in FIG. 2A, since portions of the input side edges 60CW and 60CCW (formed by the slope 49 of the elongated recess 40) in FIG. around the circumference and, consequently, hydrodynamic lubrication will be small or absent altogether, and this situation is also not optimal for preventing film leakage.

For maximum lubrication, the shape of the contact zone 16 of FIG. 2 is also more preferable than the shape of the contact zone 16 of FIG. 2B. As for the contact zone 16 shown in FIG. 2, since the first fluid 6, which wets the rotating surface 5, acts on the edge of the contact zone 16 at the edges of the contact zone re-entry 68CW or 68CCW, the first fluid 6 is most likely will fall into the contact zone 16, which circumferentially protrudes around the edges of the 66CW or 66CCW, because for the flow around the protruding around the edges of the 66CW or 66CCW, the first fluid 6 would have to move in the opposite direction of rotation. As for the contact zone 16 shown in FIG. 2B, any fluid flowing around the contact zone 16 (instead of getting into it) will still move in the direction of relative rotation. Therefore, more likely the fluid enters the contact zone 16 shown in FIG. 2, compared with the contact zone 16 shown in FIG. 2B.

As previously stated, the edges 68CW and 68CCW of re-entering the contact zone are formed as a result of compression by the rotating surface 5 of the slope 49 of the recess and the rounding 50CW and 50CCW of the recess. Preferably, the 50CW and 50CCW roundings are made by protrusions on the mold, which is subject to wear during operation. As the protrusions of the mold wear out, the shape of the contact zone 16 slowly changes from that shown in FIG. 2, and will eventually become more similar to the shape shown in FIG. 2B. The 50CW and 50CCW rounding depths are chosen to maximize the life.

the operation of the mold, so that through the mold, it would be possible for a long time to produce seals containing compressed contact zones with the necessary edges of the contact zone re-entry 68CW and 68CCW.

In FIG. 2C, for orientation, a rotating surface 5, a first edge of a contact zone 17, a second edge of a contact zone 18, a chamber 23, inlet side edges 60CW and 60CCW, 64CW and 64CCW places of highest pressure, edges 68CW and 68CCW of re-entry of the contact zone, are indicated. 70CW and 70CCW of the contact for clockwise and counterclockwise rotation and the dashed boundary line 72. FIG. 2C shows a contact zone 16, which is in all respects similar to this zone in FIG. 2, except that the minimum deviation width 85 more on figs. If the minimum inlet width 84 and the minimum deviation width 85 are equal, then more of the parts of the pressure contact zone 62CW and 62CCW that are located next to the minimum deviation width 85 will be located inside the annular seal strip 44 and therefore will be harder to lubricate. If the minimum deviation width 85 is made larger than the minimum inlet width 84, then a very small portion of the pressure contact zones 62CW and 62CCW will lie inside the annular seal strip 44, which is difficult to lubricate. If we make the minimum deviation width 85 greater than the minimum inlet width 84, then almost all parts of the pressure contact zones 62CW and 62CCW are well lubricated, since they are on the same line as the circumference with the hydrodynamic inlet portions that are formed between the interface point T73 and protruding along circles with edges 66CW or 66CCW respectively. The ratio of the minimum inlet width 84 and the minimum deviation width 85 can also be used to control the washing speed of the seal. The closer the minimum deviation width 85 to the minimum inlet width 84, the greater the flow of the first fluid 6 into the second fluid 7. The larger the minimum deviation width 85 with respect to the minimum inlet width 84, the smaller the first fluid 6 into the second fluid 7.

Figure 3 shows a partial view of an unidentified seal 2, which is configured (twisting or manufacturing) for compression in the radial direction and is designed to seal a surface rotating around the seal, having a cylindrical shape from the inside, such as the inner surface of the housing opening. Dynamic Edge 4 is Oriented

outwardly by the element, and the protruding static sealing lip 36 is an inwardly oriented element. The diameter of the dynamic edge 4 is larger than the diameter of the protruding static sealing edge 36. As shown, the first end portion 14 can be made essentially flat. A dynamic stop line of intersection 19 is indicated for orientation.

3A and 3B are partial views of an unidentified seal 2 that is configured (twisting or manufacturing) for axial compression and intended to seal a surface that rotates relative to the seal and is substantially flat (such as a shaft or housing flange), and these figures clearly illustrate that the present utility model can also be used in mechanical seals. Both the dynamic lip 4 and the protruding static sealing lip 36 are mainly axially oriented elements. 3A, the dynamic stop line of intersection 19 is positioned so that the second fluid 7 is outside relative to the dynamic edge 4, and the first fluid 6, i.e. the lubricating fluid, is inside with respect to the dynamic edge 4. In FIG. 3B, the dynamic stopping line of intersection 19 is positioned so that the second fluid 7 is inside with respect to the dynamic edge 4, and the first fluid 6, that is, the lubricating fluid, is outside with respect to the dynamic edge 4. First ends second portion 14 may be substantially cylindrical as shown in Figures 3A and 3B. In FIG. 3A, the first end portion 14 forms the inner surface of the seal, and in FIG. 3B, the first end portion 14 forms the outermost surface of the seal.

Figure 4-9 shows cross sections taken in the center of the elongated recess 40, which presents simplified embodiments of the utility model. Some elements of FIGS. 4-6 are indicated for orientation, namely: a dynamic edge 4, a first seal end portion 14, a second seal end portion 15, a dynamic stopping line of intersection 19, and a static stopping line of intersection 38.

To simplify, as shown in FIG. 4, the static sealing surface 20 of the protruding static sealing lip 36 may not necessarily be inclined, as shown in FIGS. 1A, 1B, 1D, 1E, and 1F.

For simplicity, the protruding static sealing lip 36 shown in FIGS. 1A-1F can generally be removed, and the static sealing surface 20 can simply be a peripheral sealing surface, which, in

generally, it is opposite the dynamic edge 4, as shown in FIG.

Although the seal embodiments shown in FIGS. 1-5 are compression fit seals that include a dynamic lip 4 made of one material and contain a spring element 21 that is made of another material, this technical solution is not what or a limitation of the utility model. The hydrodynamic shape of the dynamic edge 4 corresponding to the utility model can be adapted for various types of seals of rotating joints, and one or more sealing materials or other components can be used without departing from the scope of the invention.

In FIG. 6, for simplicity, the spring element 21 shown in FIGS. 1-5 has been removed, where, in other cases, there was a spring element, now there is an empty space in the form of an annular groove 22. Such a seal relates to seals with a flexible lip, as described in assigned patent US 5678829. The annular groove 22 determines that the dynamic lip 4 and the protruding static sealing lip 36 belong to a variety of flexible edges. This simplification is more suitable for applications where it is desirable to have a very low pressure in the contact zone to achieve the smallest torque required to overcome seal friction, and where eliminating the penetration of abrasive particles from the environment is not important.

The flexibility of the dynamic lip 4 somewhat reduces the pressure in the contact zone 16 between the dynamic lip 4 and the rotating surface 5 (see FIG. 1), which would occur if the seal belonged to the type of direct compression seals (such as those shown in FIG. 1 -5). Thus, flexibility helps provide less torque needed to overcome seal friction. The flexibility of the edge allows the use of a material with a relatively high modulus of elasticity, which otherwise would not be applicable for use in a homogeneous (without grooves) monolithic seal due to the high pressure in the contact zone that occurs in a homogeneous seal.

The simplified seal 2 shown in FIG. 6 may be made of any suitable sealing material, including elastomeric and rubber-like materials and various polymeric materials. However, it is preferred that the dynamic edge 4 is made of a reinforced material, such as a multilayer mesh reinforced elastomer.

Seal embodiments shown in FIGS. 7-9 are also adapted

so that they are placed in the annular groove 26 of the seal and compressed between the first component 8 of the mechanism and the rotating surface 5, as described above with respect to figure 1 (that is, compressed between the peripheral wall 11 of the groove 26 of the seal and the rotating surface 5).

7-9, the spring element 21 and the annular groove 22 of FIGS. 1-5 are excluded for simplicity, and the seal is a uniform (without grooves) compression type seal with a one-piece, one-piece housing 74, with a dynamic edge 4 and the static sealing surface 20 are on opposite sides. With the exception of the particular hydrodynamic geometrical shape of the dynamic lip 4 corresponding to the present utility model, the seals of FIGS. 7 and 8 are generally made according to US Pat. No. 5,230,520. In FIG. 7, the static sealing surface 20 has a beveled shape. For the purpose of further simplification in FIG. 8, the static sealing surface 20 is not chamfered. The beveled shape of the static sealing surface 20 of FIG. 7 in installations with a small differential pressure increases the compression ratio and changes the pressure distribution in the contact zone (compared to the seal of FIG. 8) in order to reduce the leakage of the first fluid 6 and improve the resistance to penetration of the second fluid Wednesday 7.

Figure 9 shows another simplified version in which the protruding static sealing lip 36 of Figures 7 and 8, and the static sealing surface 20 is a non-protruding outer surface of the seal 2. Except for the special hydrodynamic geometric shape of the dynamic lip 4 corresponding to the present useful models, the seal of Fig.9 is made, in General, according to US patent 4610319.

Preferably, the simplified seals of FIGS. 7-9 are made of an elastic material with a nominal durometer hardness ranging from about 70 to about 97 Shore A.

Compared to existing seals made of one material, the seals according to Figs. which improves the extrusion resistance.

10-15A show seals of rotating joints corresponding to preferred embodiments of the present utility model, wherein the seals are depicted in an uncompressed state, while for springing

a dynamic edge 4 to a rotating surface 5 (not shown) is provided with a spring element 21. For orientation, a first end portion 14 of the seal is indicated, which is directed to the first fluid 6. Also for orientation, a dynamic stop line of intersection 19 and a second fluid 7 are indicated.

10-12B, the dynamic lip 4 is made of an elastic sealing material, the spring element 21 is made in the form of a spring, the elastic modulus of which exceeds the elastic modulus of the dynamic edge, the spring element 21 is located in the annular groove 22. It is very preferable to use springs in hydrodynamic seals quality of spring elements, since their large modulus of elasticity allows the shape of the dynamic edge 4 to repeat large degrees of deflection and runout of the shaft and since the springs better resist you juice temperature residual deformation in compression compared with most elastomeric spring elements. In Fig. 10, the spring element 21 is made in the form of a cantilever spring, in Figs. 11 and 12B, the spring element 21 is made in the form of a beveled spiral spring, and in Figs. 12 and 12A, the spring element is a spring ring.

The preferred embodiment of FIG. 12 shows that, if desired, the region between the mating point T56 and the dynamic stop line of intersection 19 of the dynamic edge 4 can be tilted; 12, the shape of this region is generally conical.

10, 11, 12B and 13, the static sealing surface 20 is made of flexible sealing material. 12 and 12A, the seal 2 comprises a conventional housing 75, which is conventional for a spring-loaded spring ring seals with a lip projecting in the form of a cantilever, and a static sealing surface 20 is part of the housing 75. Typically, in such a seal design, the housing 75 is compressed in a hole the first component 8 of the mechanism (not shown), while the interference fit (sometimes supplemented by a coating such as sealant for gaskets) forms a seal between the static sealing surface 20 and the hole. The seal of FIG. 12 is configured to hold the second fluid 7, the pressure of which is higher than the pressure of the first fluid 6, and the seals of FIGS. 12A and 12B are configured to hold the first fluid 6, the pressure of which is higher than the pressure of the second fluid 7. In any case, the pressure drop exists inside the dynamic edge 4 and helps to press it against the rotating surface 5 (not shown).

The seal 2 of FIG. 13 forms the composite structure described in

US Pat. No. 5,738,358. This structure helps to control the hydrodynamic proppant by reducing the set force in the contact zone between the dynamic edge 4 and the rotating surface 5 (not shown). The dynamic edge 4 is made of a layer of the first elastic sealing material, and the spring element 21 is made of a layer of the second sealing material, the elastic modulus of which is usually less than the elastic modulus of the material from which the dynamic edge 4 is made. For example, an elastomer can be used to make the spring element 21, the durometer hardness of which is from 30 to 80 Shore A, and a sealing material (such as an elastomer or PTFE p asthik reinforced with carbon graphite), the hardness of the durometer of which exceeds 80 Shore A. Thus, the extrusion resistance of the dynamic edge 4 is regulated by its elastic modulus, and the pressure in the contact zone of the dynamic edge 4 and the rotating surface 5 (not shown) is mainly regulated , the elastic modulus of the spring element 21. This technical solution provides excellent resistance to extrusion and the relatively low torque required to overcome the friction of the seal during operation and during unfastening. With the low torque required to overcome seal friction, the operating temperature is minimized and the temperature-related deterioration of the seal is mitigated. Further, in comparison with the homogeneous structures shown in FIGS. 7-9 in the structure shown in FIG. 13, the curvature of the hydrodynamic geometric shape associated with compression is minimized. For a seal of the type shown in FIG. 13, the joint between the materials of which the dynamic edge 4 and the rotating surface 5 are made can take any suitable shape.

The greater the elastic modulus of the sealing material, the better the seal resists damage caused by high pressure extrusion. For the seals according to FIGS. 10-12B and the seals in accordance with other figures in which a spring element is present, for the best extrusion resistance, it is preferable that the dynamic lip 4 be made of a solid material with a relatively large modulus of elasticity, which resists extrusion well, such as a flexible polymer material, an elastomer with a large modulus of elasticity, such that the durometer hardness is in the range of 80-97 Shore A, or an elastomer reinforced with a mesh, fiber or metal m, or plastic with high temperature resistance characteristics, such as reinforced PTFE

carbon graphite, or PTFE with polyester or polyimide.

10, 11, 12B and 13, the seals comprise a static stopping line of intersection 38 and a protruding static sealing lip 36 that defines a static sealing surface 20 of an inclined shape. 13, the spring member 21 defines a protruding static sealing lip 36 and a static stop line of intersection 38.

On Fig shows a preferred embodiment in which a composite curve (shown by a dashed line), passing from point T78 to point T76 and preferably made in accordance with figv, adjusts the curved surface of the inclined section 48CCW (and / or inclined section 48CW ) Also, the specified composite curve allows you to deepen the elongated recess 40 to increase resistance to thermal expansion and wear of the mold. 14, a dynamic edge 4, a first end portion 14, a dynamic stop line of intersection 19 and a protruding static sealing lip 36 are indicated.

FIGS. 15 and 15A show enlarged partially darkened isometric views of a preferred embodiment of the elongated recess 40. FIG. 15 shows a full elevation for counterclockwise rotation, and FIG. 15A is an even larger view, for better understanding, showing only hydrodynamic the entrance of said lift. The inclined portion 48CCW is located along the circumference and is substantially aligned with the dynamic stop line of intersection 19 and therefore aligned with the possible directions of rotation. For orientation, a dynamic lip 4, a static stop line of intersection 38 of the seal, a first end portion of the seal 14, a slope 49 of the recess and a rounding 50CCW of the recess are indicated. As shown, the seal is configured for a radial seal, with the dynamic lip 4 generally radially inward toward the theoretical center line of the seal, and the diameter of the static stop line of intersection 38 of the seal is larger than the diameter of the dynamic stop line of intersection 19. The dynamic surface 51 of the seal can be slightly conical or substantially cylindrical as shown.

FIG. 15B is a surface map of the hydrodynamic inlet portion shown in FIG. 15A and is intended to better understand its shape. The profiles P76, P78 of the construction and the profile P55 of the pair are marked and serve as the first and second boundaries, establishing the length along the circumference of the hydrodynamic inlet section of one lift. Structural profiles P76, P78 represent the boundary of a portion of an elongated recess for counterclockwise rotation. Profile P78

the build is preferably a curve that intersects with the build profile P76 at point 97, wherein when the seal is compressed by a rotating surface, a curved inlet lateral edge 60CCW is formed (as shown in FIGS. 2-2B). The profiles P76 and P78 of the construction have different shapes and at least part of one of them is located at some distance around the circumference of the second. The build profile P76 forms a substantially convex curve defining a portion of the shape of the elongated recess 49 between point 97 and point 100. The build profile P76 preferably defines a substantially convex curve between point 100 and point 102, said curve forming a portion of the rounded shape 50CCW of the depression . The build profile P76 forms a substantially convex curve between point 102 and point 104, which curve defines a portion of the shape of the inclined portion 48CCW. The groove slope 49 is preferably smoothly mated with the rounded groove 50CCW at the point 100, and the groove rounded 50CCW is preferably smoothly mated with the inclined portion 48CCW at 102.

The build profile P76 preferably represents a line of mating or basic mating with an adjacent edge surface of an adjacent lift, which is preferably a clockwise rotation lift (not shown), which is a mirror image of the illustrated lift. In order to realize a smooth surface in the contact zone established by the hydrodynamic geometric shape of the inlet section of Figs. The build profile P78 preferably represents the line of mating or main mating with the adjacent portion of the 42CCW lift for counterclockwise rotation (see FIG. 15).

Card lines located between the building profiles P76 and P78 are parallel to the dynamic stop line of intersection 19 and map the surface of the elongated recess 40 of FIGS. 15 and 15A. Each map line between the construction profile P78 and the conjugation profile P55 is a primary curve 52, the end points of which are preferably smoothly mated to the adjacent geometric shape of the edge surface (i.e., at the mating points T78 and T55). Each map line between the construction profile P76 and the conjugation point T55 is a secondary curve 54, the endpoints of which are preferably smoothly mated with the adjacent geometric shape of the edge surface (i.e., at the mating points T76 and T55). All different primary curves 52 display

the primary curved surface bounded by the pairing profile P55 and the build profile P78. All of the various secondary curves 54 represent the secondary curved surface bounded by the mating profile P55 and the construction profile P76. Preferably, at least a portion of the aforementioned primary curved surface smoothly mates or substantially seamlessly mates with the aforementioned secondary curved surface along the mating profile P55.

The pairing profile P55 represents the pairing line of the primary curve 52 and the secondary curve 54 of each map line. As shown in FIG. 15B, the circumferential arrangement of the mating profile P55 is preferably not constant for at least a portion of the length of the mating profile P55.

One example of the map line construction method of FIG. 15B is shown in FIGS. 16 and 16A; this method automatically adjusts the curvature of the primary curve 52 as a function of the diameter of the seal, which provides a narrowing to the shaft for a family of seals of different diameters, while the sizes of the seals are selected for various external dimensions of the equipment.

The method for constructing the radiuses of mates, shown in FIG. 15B and shown in more detail in FIGS. 16 and 16A, allows two different profiles to be smoothly mated in a circumferential direction, such as, for example, the construction profiles P76 and P78 of FIG. 15B, which provides a hydrodynamic geometric shape of the input section. By changing the shapes of the connected construction profiles, you can get a number of different performance characteristics.

Figs. 14-15B show seals intended to be sealed on the outside of a cylindrical surface such as a shaft, however, the main cross-sectional shapes are suitable (after twisting or during manufacture) for mechanical seals or for seals from the inside of a cylindrical surface, exactly as described in the description seals according to figures 1-1F.

16-16C schematically shows a construction method that can be used to obtain map lines in the area limited by the construction profiles P76 and P78 of FIG. 15B, which define the surface of the elongated recess 40 shown in FIGS. 15 and 15A. The construction methods of FIGS. 16-16C can be used to connect any two construction profiles to form a hydrodynamic inlet section, regardless of whether one of the construction profiles touches the circumference of the other profile. Also, the construction method shown in Fig.16-16C, can be used to establish composite curvature

1C to FIG. 1F — in this case, it is better to use the perspective view of the inclined section 48CW shown in FIG. 1B. In any case, it is preferable that the construction method lie in a plane of triangular projection located at right angles to the theoretical axis a seal designed to seal the shaft rotating relative to it (i.e., parallel to the first end portion 14 of the seal and the dynamic stop line of the suppression 19, which are shown in figa, 15 or 15A).

The construction lines 86, 88 shown for reference are radially spaced lines starting from the theoretical center line / seal axis 90 (see FIG. 16A). The construction line 88 of FIGS. 16-16C is located on the section plane D-D of FIG. 1B. The theoretical arc 92 represents a line that is smoothly conjugated with the primary curve 52 and extends from the construction line 86 to the construction line 88. Theoretical curve 92 represents the diameter of the seal surface adjacent to the primary curve 52 (see, for example, FIGS. 1A and 15). The primary curve 52 is preferably smoothly mated with the theoretical arc 92 and, therefore, smoothly mated with the adjacent seal material, which is the theoretical curve 92. The primary curve 52 is preferably smoothly mated with the secondary curve 54 at the mating point T55.

Secondary curve 54 is designed so that the 42CW rise for clockwise rotation and the 42CCW rise for counterclockwise rotation (see FIG. 1A) do not intersect at an acute angle and to provide a recess width 96 that is designed to accommodate various tolerances, changes in the amount of compression and changes in the size of the seal during temperature changes during operation of the seal, to prevent general smoothing of the geometric shape of the seal relative to the rotating surface 5. It is preferable that there is no the aforementioned intersections at an acute angle, since sharp angles when intersecting rotating seals in seals sometimes lead to fatigue damage to the rubber and since the geometrical shape of the mold, due to which sharp angles are formed when crossing, wears out quickly and therefore the dimensions of the seal can increase and with large values of thermal expansion, swelling of the seal associated with the environment, the axial load of the seal or its compression, its performance may deteriorate.

The primary curve 52 sets the degree of narrowing of the inclined portion 48CW and / or the rotating surface 5 when viewed in a circumferential direction (see FIGS. 1 and 17). Narrowing speed is a crucial aspect of hydrodynamic

performance, since it controls the rate of increase in pressure in the direction along the circumference in the contact zone (known as DPDX) between the dynamic edge 4 and the surface 5 rotating relative to it.

For example, if the radius of the primary curve 52 remains constant for a family of radial seals of different diameters that seal the shaft (in order to correspond to the different external dimensions of the equipment, as described above with respect to FIG. 17), then the speed of narrowing with the rotating surface 5 of the shaft changes as function of the diameter of the rotating surface 5. For a given radius of the primary curve 52, the larger the shaft, the smoother the narrowing speed, while the primary curve 52 has a more efficient shape for the hydrodynamic nical wedge. On the contrary, the smaller the diameter of the shaft for a given radius of the primary curve 52, the less smooth the narrowing speed, which makes the shape of the primary curve 52 less effective for maintaining a hydrodynamic wedge. The use of constant curvature in the entire family of seals of different diameters, leading to different constrictions for different seals of the family, is a disadvantage of the existing prior art, which is described in connection with FIG. In a preferred embodiment of the present invention, the constriction rate is approximately the same for each seal in the family of seals of different diameters, so that the hydrodynamic shape maintains the constant efficiency of the hydrodynamic wedge for the entire family of seals of different diameters. Thus, all seals of a family of seals of different diameters have the same hydrodynamic performance. This principle is described in detail with reference to FIGS. 17 and 17A.

With the geometric shape shown in FIGS. 16-16C, the degree of narrowing of the rotating surface 5 and either the hydrodynamic inlet portion of FIGS. 15-15B or the inclined sections 48CW, 48CCW of FIGS. 1-1F can be maintained substantially constant for families of radial seals with different diameters if the width 94 of the primary arc and the length 80 of the primary arc of the primary curve 52 are kept constant. The length 80 of the primary arc is not an angle, but represents the length of the arc, which is measured along the theoretical arc 92 from the mating point T78 to point 79. Similarly, the length 82 of the secondary arc is not an angle, but rather the length of the arc, which is measured along the theoretical arc 92 from point 79 to point 81. The total length 83 of the arc is the sum of the length 80 of the primary arc and the length 82 of the secondary arc.

Using the construction method shown schematically in FIGS. 16 and 16A,

The preferred way to maintain a substantially constant degree of narrowing in the family of radial seals of different diameters with the same recess width 96 is to maintain a substantially constant length 83 of the composite arc for the family of seals of different diameters, while also maintaining a constant ratio of the length 82 of the secondary arc to a length of 83 of the composite arc. If the width 96 of the recess varies from seal to seal, then it is necessary to adjust the shape of the primary curve 52 to maintain the same narrowing for all seals.

With the aforementioned restrictions on the width 94 of the primary arc and the length 80 of the primary arc (to achieve close narrowing values for all seals in the family), the bottleneck that needs to be considered is that if the primary curve 52 is located so that it is smoothly connected to the theoretical arc 92, the primary curve 52 (and the inclined portion 48CW) has a substantially convex shape in large diameter seals (as shown in FIG. 16) and has a substantially concave shape in small diameter seals (as shown in FIG. 16A ) and, the investigator on supports substantially the same speed restriction inclined section 48CW and rotating it relative to the surface 5 for the entire family of radial seals of different diameters regardless of the circumferential length of the lift. Therefore, for a given curvature, there may be values of the diameter of the seal, for which the primary curve 52 will be a straight line (a curve with an infinite radius), seals with a large diameter will have, in general, a convex primary curve 52; and seals with a smaller diameter will have, in general, a concave primary curve 52. With the concave shape of the primary curve 52, the rounding of the recess also has a concave shape when looking at the cross section in the circumferential direction. Therefore, at least a portion of the hydrodynamic inlet portion formed by the elongated recess will have a concave shape when viewed from a cross section in a circumferential direction.

For example, if a family of seals is made using a patterned seal with an inner diameter of 2.75 "and a convex primary curve 52 with a radius of 0.50", then the primary curve 52 for a seal with an inner diameter of 0.6 "will have a radius of approximately 1.468", and the primary curve 52 for a seal with a diameter of 16.5 "it will be concave and have a radius of approximately equal to 0.383". The primary curve 52 of all seals from the above family will have essentially the same degree of narrowing with the corresponding shafts, even if their actual curvature will differ and even if some of them will be concave and others convex.

Returning to FIG. 15B, it is obvious that the change in position of the conjugation profile P55 is observed because the ratio of the length of the secondary arc 82 to the length 83 of the composite arc changes in a sinusoid from one line of the card to another, while the ratio changes from 0.5 at point 97, where the profiles P76 and P78 of the construction relate, to a smaller ratio at point 98, where the profile P55 becomes almost parallel to the profile P76. This sinusoidal change in ratio helps to obtain a non-faceted, smooth contour rounding 50CCW of the recess, which is shown in FIGS. 15 and 15A.

For the design of the elongated recess shown in FIGS. 15 and 15A and made by the map lines of FIG. 15B, the primary curve 52 is generally convex in some places and generally concave in other places, depending on the length 80 of the primary arc and width 94 of the primary arc. For example, a part of the primary curves 52, which are located near point 97, are clearly concave, and those primary curves 52, which are further from point 97, are clearly convex.

Thus, in FIG. 15B, the exact narrowing with the shaft changes from one line of the card to the next. It is important that the narrowing from one line to another is repeated when moving from a seal of one diameter to a seal of another diameter, provided that the length 80 of the primary arc and the width 94 of the primary arc remain constant or provided that the length 83 of the composite arc and the ratio of the length 82 of the secondary arc and the length of the 83 composite arcs remains constant for any given map line among the entire family of seals of various diameters.

Returning again to FIGS. 1-1F. The narrowing of the elongated recess 40 and the rotating surface 5 in the circumferential direction is controlled by three independent elements: the inclined sections 48CW and 48CCW, the slope 49 of the recess and the curves 50CW and 50CCW of the recess. A method for controlling the narrowing of the rotating surface 5 and the inclined portions 48CW and 48CCW for a family of seals with different diameters has already been described above. Assuming that the narrowing of the inclined sections is also the same, the narrowing along the circumference between the rotating surface 5 and the slope 49 of the recess of Figs. 49 recesses are the same for the entire family of seals with different diameters and holds for the same distance along the circumference. The narrowing of the rotating surface 5 and the curves 50CW and 50CCW of the recess of FIGS. 1-1F can be supported for the family of seals with

different diameters, if the change in the radius of curvature of 50CW, 50CCW is the same for a family of seals with different diameters and takes place for the same distance along the circumference.

Without going beyond the scope of the utility model, the primary curve 52 and the secondary curve 54 can take any suitable shape, such as part of a circle, part of an ellipse, part of a sinusoidal curve, part of a parabolic curve, part of a cycloid, part of a persper / version, a combination of curves, a combination of curves and straight lines (that is, curves of infinite radius of curvature) and so on.

For example, in FIG. 16B, the secondary curve 54 is shown as an S-shaped curve, which makes it easier to obtain a larger recess width 96, and in FIGS. 16 and 16A, the secondary curve 54 is shown as a circular arc. On Fig.16-16C, the primary curve 52 is shown in the form of an arc of a circle, which is shown in Fig.16-16C having the same size. The center of the primary curve 52 lies along the construction line 86 in FIGS. 16-16C. The center of the secondary curve 54 is located along the construction line 88 in FIGS. 16-16A. Thus, said construction line 86 and construction line 88 are straight lines of infinite length, and they extend beyond the boundaries of the drawing.

The construction method shown in FIG. 16B was used to create an elongated recess 40 shown in FIG. By means of the construction method shown in FIG. 16B, a recess of any desired width 96 can be created.

In Fig. 16C, the shape of the secondary curve 54 can optionally be used to distance the primary curve 52 from the construction line 88, however, it should be noted that adjacent lifts formed using this method will still adjoin each other by means of a slope 49 of the recess (slope 49 depressions shown in other Fig., For example, figa).

The construction method shown in FIG. 16D can be used to establish the degree of change in the variable curvature of the slope 49 of the recess (such as that shown in FIGS. 1-1F) and, therefore, to regulate the narrowing of the slope 49 of the recess and the rotating surface 5, and thus thus, it is used to control an important factor that affects the narrowing of the rotating surface 5 and the 50CW and 50CCW roundings.

It is preferable that the construction method according to FIG. 16D is carried out on a plane of triangular projection, which is located at right angles to the theoretical axis of the seal, designed to seal a rotating shaft (i.e. parallel to the first end portion 14 of the seal and the dynamic retaining

suppression lines 19, which are shown in FIGS. 15 and 15A).

In Fig. 16D, some elements used in the description of Fig. 16 are indicated, namely: primary curve 52, secondary curve 54, conjugation point T55, conjugation point T76, conjugation point T78, point 81, composite curve length 83, construction line 86, construction line 88 and theoretical arc 92. In addition to these elements, FIG. 16D also shows a primary slope curve 152, which is substantially smoothly conjugated with a secondary slope curve 154 at mating point T155 and smoothly mated with theoretical arc 92 and dynamic seal surface 51 at point T17 8 pairings. The primary slope curve 152 and the secondary slope curve 154 are the theoretical intersection points of the multifunctional edges 46CW and 46CCW and the recess slope 49 (as shown in FIGS. 1-1F, the 50CW and 50CCW roundings of the recess cover these theoretical intersections). The theoretical intersection points mentioned above are located where the multifunctional edges 46CW and 46CCW and the slope 49 of the recess intersect if their curved surfaces are continued until they intersect at an acute angle, in the absence of curvature 50CW and 50CCW of the recess.

The construction line 186 shown for reference is a radial line starting from a theoretical center line / axis of compaction. The degree of narrowing of the slope 49 of the recess and the rotating surface 5 can be kept substantially constant for a family of radial seals with different diameters, if the width 194 and the length 180 of the primary arc 152 of the slope are kept constant. The length of the arc 180 is not an angle and is measured along the theoretical arc 92 from the point T178 of the interface to point 179. Similarly, the length 182 of the arc is not an angle and is measured along the theoretical arc 92 from point 179 to point 81. The total length of the arc 183 is the sum of the length of 180 of the arc and lengths of 182 arcs. Typically, the length 183 of the composite arc is longer than the length 83 of the composite arc. A composite arc with a length of 183 and / or a composite arc with a length of 83, in a preferred embodiment, can reach the lower elevation point, which is located on the section plane F-F of FIG. 1B.

In fact, using the construction method shown schematically in FIG. 16D, the preferred way to maintain a substantially constant narrowing speed of the slope 49 of the recess in the family of radial seals of different diameters with the same curve width 196 is to maintain a substantially constant length 183 a compound arc for a family of seals of different diameters, while also maintaining a constant ratio of the length of the arc 182 to the length 183 of the composite arc. If the width of the curve 196 varies from seal to seal, then to maintain the same narrowing for all seals, you must accordingly

adjust the shape of the primary slope curve 152.

With the aforementioned restrictions on the width of the arc 194 and the length of the arc 180 (to achieve close narrowing values for all seals of the family), the bottleneck that needs to be considered is that if the primary slope curve 152 is located so that it is smoothly connected with the theoretical arc 92 , the primary slope curve 152 has a substantially convex shape in large diameter seals (as shown in FIG. 16D) and has a substantially concave shape in small diameter seals and therefore, substantially the same speed is maintained narrowing slope recess 49 and rotating it relative to the surface 5 for the entire family of radial seals of different diameters regardless of the circumferential length of the lift. Therefore, for a given curvature, there may be a value for the diameter of the seal, for which the primary slope curve 152 will be a straight line (a curve with an infinite radius), seals with a larger diameter will have, in general, a convex primary slope curve 152, and seals with a smaller diameter will have generally a concave primary slope curve 152.

Without going beyond the scope of the utility model, the primary slope curve 152 and the secondary slope curve 154 can take any suitable shape, such as part of a circle, part of an ellipse, part of a sinusoidal curve, part of a parabolic curve, part of a cycloid, part of a versper / versioner, a combination of curves, a combination of curves and straight lines (that is, curves of infinite radius of curvature) and so on. For example, the shape of the secondary slope curve 154 may substantially coincide with the shape of the secondary curve 54 shown in FIG. 16B or 16C.

The cross-sectional width of the seal is the distance between the static sealing surface 20 and the dynamic sealing surface 51, which is usually measured between the dynamic stopping line of intersection 19 and the static stopping line of intersection 38 with an uncompressed seal 2. It was previously noted that for the same compression values, seals installed with interference fit, with a smaller cross-sectional width, have a greater compression, expressed as a percentage (compared with seals with a larger cross-sectional width), Therefore, a much higher pressure in the contact zone, which complicates lubrication seals with a small cross section. For example, for a seal with a cross-sectional width of 0.212 "and a compression of 0.023", the pressure in the contact zone is approximately 52% greater than for a seal with a cross-sectional size of 0.335 "and a compression of 0.026". It was also noted earlier that the hydrodynamic inlet section, which maintains, in general, the same

the degree of narrowing for a family of seals of various diameters is an improvement on the state of the art. In addition, a significant improvement in the existing level of technology is to reduce the rate of narrowing of the hydrodynamic inlet section and shaft (i.e., the narrowing has become smoother) for seals with a small cross section and to provide a stronger hydrodynamic proppant to compensate for the increase in pressure in the contact zone that occurs for small cross-sectional seals. Thus, for a family of seals containing seals of different cross-sectional widths for various external dimensions of the equipment, it is preferable that the degree of narrowing of the hydrodynamic inlet portion be smoother to seal a small cross-section to compensate for the increase in pressure in the contact zone caused by the smaller cross-sectional width to provide the same lubrication level for a family of seals with different cross-sectional widths.

Based on the foregoing, in this utility model, all tasks are achieved, and it has all the features indicated above, and there are other useful tasks and features that are inherent in the device described here.

Since some specific hydrodynamic seals of rotating joints and the shapes of the sealing elements are described, however, a large number of other geometric shapes are possible, using the basic ideas and principles of this utility model.

The above description of the utility model illustrates and explains the present utility model. Without departing from the essence of the utility model, one can propose various changes in sizes, shapes and materials, as well as details of the illustrated construction. Therefore, the present embodiment of the utility model should be considered as an illustration, and not a limitation, the scope of the utility model is limited by the utility model formula, and not the description given, all changes that are equivalent to the utility model formula clauses are covered by the utility model formula.

Claims (56)

1. The hydrodynamic seal of the rotating joint, located next to the component of the mechanism for dynamically sealing the surface rotating relative to it and made in the form of a partition between the first fluid from the lubricant source and the second fluid, containing a static sealing surface, forming a sealing joint with a component of the mechanism; basically, an annular dynamic edge, wherein at least a portion of the dynamic edge is in a compressed state in contact with a rotating surface and forms a contact zone at the interface between the dynamic edge and the rotating surface, while during rotation of the dynamic edge and the rotating surface relative to each other, of the contact is a moving sealed joint in which the dynamic edge glides relative to a rotating surface, and in the absence of relative rotation contact she is a static sealed joint; wherein said dynamic edge comprises: a first edge side directed toward the first fluid; a second side of the edge directed toward the second fluid; a first lift comprising a first multifunctional edge located adjacent to the first side of the edge, wherein at least a portion of the first multifunctional edge is beveled relative to the direction of relative rotation; as well as a second lift containing a second multifunctional edge located next to the first side of the edge, and at least part of the second multifunctional edge is beveled relative to the direction of relative rotation and beveled in the opposite direction relative to the beveled part of the first multifunctional edge; and an elongated recess defining a first hydrodynamic inlet portion for a first lift and a second hydrodynamic inlet portion for a second lift. 2. The seal according to claim 1, in which the beveled multifunctional edges form contact zones with increased pressure inside the contact zone. 3. The seal according to claim 1, in which the first and second lifts are made, essentially, mirror images of each other. 4. The seal according to claim 1, in which the first lift is made in the form of a lift for rotation clockwise, and the second lift is made in the form of a lift for rotation counterclockwise, and the lift for rotation clockwise is designed to deflect the film of the first fluid in the direction of the second side of the edge with a relative clockwise rotation, and the rise for counterclockwise rotation is designed to deflect the film of the first fluid in the direction of the second side of the edge with a relative rotation counterclockwise arrow. 5. The seal according to claim 1, in which, essentially, the first and second multifunctional edges are beveled relative to the direction of relative rotation. 6. The seal according to claim 1, in which the multifunctional edges are curved, and the curvature of at least part of the multifunctional edges varies along at least a portion of the length of the multifunctional edges. 7. The seal according to claim 1, in which the curvature of the multifunctional edges changes sinusoidally. 8. The seal according to claim 1, which contains a spring element for springing the dynamic edge relative to the rotating surface. 9. The seal of claim 8, in which the spring element is made in the form of an elastomeric ring. 10. The seal of claim 8, in which the spring element is made in the form of a cantilever spring. 11. The seal of claim 8, in which the spring element is made in the form of a spring ring. 12. The seal of claim 8, in which the spring element is made in the form of a beveled cylindrical spring. 13. The seal of claim 8, in which the dynamic edge is made of elastic material, and the spring element is made in the form of a spring, the elastic modulus of which exceeds the elastic modulus of the specified dynamic edge. 14. The seal of claim 8, in which the dynamic edge is made of a layer of the first sealing material, and said spring element is made of a second sealing material. 15. The seal according to claim 1, in which the static sealing surface is made as part of a protruding static sealing edge. 16. The seal according to claim 1, in which the static sealing surface is made inclined. 17. The seal according to claim 1, which contains a monolithic seal of uniform cross section for sealing the shaft in the radial direction. 18. The seal according to claim 1, in which the contact zone contains the first and second input side edges of a curved shape, while the input side edges are beveled relative to the direction of relative rotation of the dynamic edge and the rotating surface. 19. The seal according to claim 1, in which the contact zone defines a minimum entrance width and a minimum deviation width, said minimum entrance width being less than the minimum deviation width. 20. The seal of claim 18, wherein, in a clockwise direction, the beveled first inlet lateral edge is configured to function as a hydrodynamic inlet portion, and in a counterclockwise direction, the beveled second inlet lateral edge is configured to function as a hydrodynamic inlet portion . 21. The seal of claim 18, wherein the first and second inlet side edges are adjacent to each other. 22. A hydrodynamic seal of a rotating joint located next to a component of a mechanism for dynamically sealing a surface rotating relative to it and made in the form of a partition between a first fluid from a lubricant source and a second fluid, comprising: a static sealing surface forming a sealing joint with a component of the mechanism; a generally annular dynamic edge comprising: a first lift comprising a first multifunctional edge, wherein at least a portion of the first multifunctional edge forms a first inclined portion; a second lift comprising a second multifunctional edge, wherein at least a portion of the second multifunctional edge forms a second inclined portion; moreover, the first and second inclined sections are curved towards the rotating surface; as well as an elongated recess containing a slope of the recess, the profile of which in the direction perpendicular to the cross section is mainly a convex curve, with the first side of the elongated recess partially formed by the first inclined section, and the second side partially formed by the second inclined section; both the first and second rounding of the recess, wherein the profile of the rounding of the recess when viewed perpendicular to the cross-section, is basically a concave curve, and in the direction along the circumference of the rounding of the recess have an inclined profile, and these roundings are essentially smoothly mated with the slope of the recess . 23. The seal according to item 22, in which the curvature of the specified slope of the recess varies. 24. The seal according to item 22, in which the curvature of at least one of the specified first and second rounding of the recess. 25. The seal according to item 22, in which each inclined section regulates the established narrowing of the elongated recess and the rotating surface. 26. The seal according to item 23, in which the degree of change of the variable curvature of the slope of the recess is partially regulated by the established narrowing of the slope of the recess and the rotating surface. 27. The seal according to paragraph 24, in which the degree and magnitude of the change in the variable curvature of at least one of the slopes of the recess is partially regulated by the established narrowing of the slope of the recess and the rotating surface. 28. The seal according to item 22, in which at least part of each of these inclined sections is compressed and touches the rotating surface and at least part of each of these inclined sections is inclined from the rotating surface, and its shape contributes to the formation of a hydrodynamic proppant actions. 29. The seal according to item 22, in which the dynamic edge forms the first edge of the re-entry of the contact zone by compressing the slope of the recess and the first rounding of the recess by the rotating surface and forms the second edge of the re-entry of the contact zone by compressing the slope of the recess and the second rounding of the recess by the rotating surface. 30. Hydrodynamic seal of a rotating joint, located next to a component of the mechanism for dynamic sealing of the rotating surface and made in the form of a partition between the first fluid from the source of lubricant and the second fluid, containing: a static sealing surface forming a sealing joint with a component of the mechanism; basically, an annular dynamic edge, wherein at least a portion of the dynamic edge in a compressed state is in contact with the rotating surface and forms a contact zone in the interface between the dynamic edge and the rotating surface, and the contact zone comprises a first non-annular edge of the contact zone directed towards the first fluid, and contains a second edge of the contact zone directed to the second fluid, the first edge of the contact zone contains a first protruding along the circumference of the edge, which along the circumference in It extends beyond the boundaries of the first edge of the re-entry of the contact zone. 31. The seal of claim 30, wherein the first edge of the contact zone comprises a second edge protruding along the circumference, which protrudes along the circumference of the second edge of the reentry of the contact zone and is opposite the first edge protruding along the circumference. 32. The seal of claim 31, wherein the second edge of the contact zone is substantially annular in shape and substantially aligned with the direction of rotation between the dynamic edge and the rotating surface. 33. The seal according to p, in which the specified dynamic edge contains a dynamic stop line of intersection of a discontinuous annular shape, while the line of intersection is mainly aligned with the direction of relative rotation of the dynamic edge and the rotating surface. 34. A hydrodynamic seal of a rotating joint comprising a dynamic edge defining at least one hydrodynamic inlet portion delimited by adjacent edge surfaces and first and second boundaries establishing an arrangement along the circumference of the hydrodynamic inlet portion, wherein the first and second boundaries have different profiles and at least parts of the first and second boundaries are located at a certain distance along the circumference from one another, and at least part of the input the line is essentially smoothly interfaced with the edge surfaces at the first and second boundaries, and at least a portion of the hydrodynamic inlet portion is formed by a primary curved surface and a secondary curved surface, with at least a portion of the primary curved surface paired with a secondary curved surface. 35. The seal of claim 34, wherein at least one of said first and second boundaries is bent along a circle toward another boundary. 36. The seal according to clause 34, in which one of the first and second boundaries forms two basically convex curves connected by a curve, basically concave in shape. 37. A system of compression hydrodynamic seals for rotating joints designed to hold a fluid, the seals of this type being arranged to be compressed by a rotating surface, wherein in the installed position, said seals form a contact zone and pressure in the contact zone with the rotating surface, containing many separate seals , the sizes of which correspond to various external dimensions of the equipment, and the cross-sectional dimensions of some of these are separate x seals smaller than the cross-sectional dimensions of other seals; while in the installed position, each of these individual seals contains at least one hydrodynamic inlet section that sets the degree of narrowing with a rotating surface and forms a hydrodynamic proppant with relative rotation, while the fluid film is wedged into the contact zone; moreover, the pressure in the contact zone of the individual seals varies as a function of the size of the cross section, and the pressure in the contact zone of the individual seals with a smaller cross section exceeds the pressure in the contact zone of these individual seals with a large cross section; moreover, the degree of narrowing is less for these individual seals with a smaller cross-section than the degree of narrowing for these individual seals with a large cross-section to compensate for the greater pressure in the contact zone of these individual seals with a smaller cross-section. 38. A system of hydrodynamic seals for rotating joints designed to hold a fluid, the seals of this type being arranged to be compressed by a rotating surface, wherein in the installed position, said seals form a contact zone and pressure in the contact zone relative to the rotating surface, containing a plurality of separate seals, the sizes of which correspond to various external dimensions of the equipment, and the diameter of some of these individual seals is smaller than the diameter other seals; while in the installed position, each of these individual seals contains at least one hydrodynamic inlet section that sets the degree of narrowing with a rotating surface and forms a hydrodynamic proppant with relative rotation, while the fluid film wedges into the contact zone; moreover, the degree of narrowing for these individual seals of smaller diameter, essentially coincides with the degree of narrowing for these individual seals with large diameter. 39. The system of clause 38, in which the degree of narrowing is partially controlled by individual seals containing primary curves that determine the length and width of the primary arc, while at least some of these individual seals have the same length of the primary arc and the width of the primary arcs. 40. A system of hydrodynamic seals for rotating joints designed to hold a fluid, the seals of the indicated type being adapted to be compressed by a rotating surface, wherein in the installed position, said seals form a contact zone and pressure in the contact zone relative to the rotating surface, comprising a plurality of individual seals, the sizes of which correspond to various external dimensions of the equipment, and the diameter of some of these individual seals is smaller than the diameter other seals; while in the installed position, each of these individual seals contains at least one hydrodynamic inlet section having a certain bend and setting the degree of narrowing with a rotating surface and forming a hydrodynamic proppant with relative rotation, while the fluid film wedges into the contact zone ; moreover, at least a portion of the bend of said individual seals of a smaller diameter is a concave curve, and at least a portion of said bend of said individual seals of a larger diameter is a convex curve. 41. A hydrodynamic seal of a rotating joint, the edge of which is designed to hold a fluid that is compressed in a fixed position by a rotating surface, and in an unspecified position, the sealing lip contains at least one hydrodynamic inlet portion comprising a bend in a circumferential direction, and, at least a portion of said bend is concave. 42. A radial hydrodynamic seal of a rotating joint, designed to hold a fluid that is compressed in a fixed position by a rotating surface, and in an unspecified position, a radial seal contains a dynamic edge with at least one hydrodynamic inlet section, and at least a portion of said hydrodynamic inlet portion, which is determined by a rectilinear element substantially spaced along the circumference. 43. Hydrodynamic sealing of a rotating joint, the hydrodynamic inlet portion of which is designed to seal a rotating joint, formed at least partially by the first and second surfaces, mainly convex in shape, moreover, these surfaces are essentially smoothly conjugated with, basically, concave a surface located between said first and second surfaces is generally convex in shape. 44. The seal according to item 43, in which the curvature of at least one of these first and second surfaces, mainly convex in shape. 45. The seal according to item 43, in which the curvature of the specified concave surface is variable. 46. The seal according to item 43, in which one of the first and second surfaces, mainly convex in shape, is made in the form of an inclined section, curved so that it touches a rotating surface that is at least partially open to the lubricant. 47. The seal of claim 46, wherein the rotating surface is rotatable in two directions relative to said first and second surfaces of substantially convex shape, the inclined portion being part of a multifunctional edge that deflects the lubricant in one of two possible relative rotation directions and hydrodynamically wedges the lubricant into a sealed joint in the other of the two possible directions of relative rotation. 48. The seal of claim 46, wherein at least a portion of the first and second substantially convex surfaces and at least a portion of the substantially concave surface is in contact with the rotating surface and forms at least a portion contact zones. 49. The seal of claim 48, wherein the other of said first and second generally convex surfaces is made in the form of a lateral part of the inlet, by means of which said contact zone comprises a beveled inlet lateral edge. 50. The seal according to claim 48, wherein by means of a substantially concave surface in the contact zone, an edge of re-entry of the contact zone is formed and by an inclined portion in said contact zone, an edge protruding along the circumference is formed which extends beyond said edge of re-entry of the contact zone. 51. The seal according to item 46, in which the first and second, mainly convex surfaces and mainly concave surface together form at least part of the elongated recess; and the bend of the specified inclined section in contact with the rotating surface is set by the primary curve, which is essentially smoothly conjugated with the secondary curve, which sets the width of the recess in this place, which provides clearance with the rotating surface in adverse operating conditions. 52. The seal of claim 51, wherein the secondary curve is an S-shaped curve. 53. The hydrodynamic seal of a rotating joint, the hydrodynamic inlet portion of which contains a primary curved surface that defines at least a portion of the elongated recess, while the primary curved surface contains a convex part, when viewed from the cross section along the circumference. 54. The seal of claim 53, wherein the primary curved surface is substantially smoothly mated to the secondary curved surface at the interface and at least a portion of said secondary curved surface is concave when viewed from a cross section along a circle. 55. The seal according to item 53, which provides for the ability to change the location of the specified interface in the direction along the circumference. 56. The seal according to item 53, which provides for the ability to change the location of the specified interface in a sinusoid in the direction along the circumference.