RU57334U1 - HYDRODYNAMIC LUBRICATED SEAL OF A ROTATING SHAFT HAVING A GEOMETRY RESISTANT TO TURNING - Google Patents

Abstract

A hydrodynamically lubricated gland type seal of a rotating shaft with a profile geometry suitable for holding pressurized lubricant is proposed, which in a preferred embodiment includes a convex static sealing surface, which, compared with the prior art, dramatically improves the protective effect of the dynamic sealing interface in low pressure applications and without pressure, by ensuring symmetrical deformation of the seal on static and dynamic plotnitelnyh the interfaces. Under abrasive conditions, an improved protective effect leads to a sharp decrease in seal and shaft wear compared with the prior art and provides a significant increase in service life. The invention also extends the service life by enabling higher levels of initial compression compared to the prior art without impairing hydrodynamic lubrication; this additional compression makes the seal less sensitive to permanent compression deformation, abrasion, mechanical misalignment, dynamic runout and manufacturing tolerances, and also makes hydrodynamic seals with smaller profiles more practical. Figure 3

Description

FIELD OF THE INVENTION

This invention relates generally to hydrodynamically lubricated rotary shaft seals and, in particular, to a hydrodynamically lubricated rotary shaft seal having a convex geometry of a static sealing interface that minimizes the possibility of twisting of the seal, which otherwise may result from radial compression.

State of the art

In the oil industry, hydrodynamically lubricated, rotary shaft packing rings, which implement the principles set forth in US Pat. No. 4,610,319 and sold by Kalsi Engineering, Inc., are widely used. of Sugarland, Tx. under the brand name Kalsi Seals®. Hydrodynamically lubricated seals are used to maintain lubricant retention and eliminate contamination in harsh abrasive conditions, such as well drilling conditions in an oil field, and have been successfully used in installations with both low and high pressure lubricants. Commercial applications related to the oilfield include rotating conical bits, hydraulic downhole motors, high-speed core swivels and rotating drill heads. All references herein to hydrodynamically lubricated seals or hydrodynamic seals refer to seals that implement the principles of the above US patent.

11-18 of this description shows hydrodynamic seals according to the prior art, which are explained here to facilitate understanding of the differences between these seals according to the prior art and seals according to this invention. One of the first applications of the hydrodynamic sealing of a rotating shaft in the oil industry is the use of low-pressure lubricants, namely in a rotating conical drill bit, which is shown as an example in FIG. 11. The geometry of the hydrodynamic seal profile for a rotating conical drill bit is determined by the shape, length and relative position of the shaft interface and the relative

seal groove position. This leads to dynamic contact geometry, which is located approximately in the center of the static seal geometry. The seal of the drill bit defines a static sealing surface 1 of the section, side 2 of the environment for contact with the drilling fluid and side 3 of the lubricant for contact with the lubricant in the lubrication chamber of the bit. The seal also forms a sharp protective edge 4 and a dynamic sealing sealing surface 5 of the section, which are in contact with the cylindrical sealing surface 6 of the shaft and form a non-circular hydrodynamic edge 7, which is open for lubrication. The laterally shifted ratio of the flat surface 2 on the seal or mud drilling side to the sharp circular protective edge 4 is dictated by the radius of curvature of the shaft 8. This feature causes the dynamic sealing surface of the drill bit seal interface to be substantially central relative to the static seal interface . Therefore, the radial compression of the hydrodynamic seal of a rotating drill bit is not prone to cause twisting of the seal. The development of the cylindrical contour of the sealing surface of the drill bit sealing interface is shown in FIG. 13.

Drill Bit Seal Testing

Extensive field tests, which include samples of hydrodynamic seals of the drill bit and conventional non-hydrodynamic seals of the oval-shaped bit, showed that the hydrodynamic and protective properties of the hydrodynamic seals effectively influenced the reduction of wear on the seal and the mating surface of the rotating shaft. In some cases, the inner surfaces of the hydrodynamic seals were truly wear-free after a lengthy test; surface finish was still visible. Under the same operating conditions, standard non-hydrodynamic seals, which were used as control samples, usually showed a significant formation of grooves on the inner surface, and also led to the formation of grooves on the shaft surface associated with the seal. Based on an analysis of the field test data, the drill manufacturer concluded that “the overall service life and reliability of a sealed cone bit with a sliding bearing are significantly improved when using a hydrodynamically lubricated seal,” and the seal has been widely used in rotating conical drill bits.

The indicated geometry of the hydrodynamic compaction for the drill bit was subjected to extensive finite element analysis during its development with

the purpose of controlling the shape and position of the dynamic sealing surface of the interface, the nature of the contact pressure profile of the dynamic interface, and placing the full profile inside the packing box. The thickness of the drill seal film and the predicted contact pressure between the surfaces show that the lubrication caused by the wedge-shaped action is adequate under normal operating conditions, and also shows that the pressure distribution between the surfaces is very desirable since it rises sharply near the protective edge.

The shape of the contact zone between the seal surfaces of the drill was also carefully experimentally measured under various pressure and temperature conditions that are encountered under real conditions. The results showed that although the contact zone between the drill seal surfaces or the outline indicated by F in FIG. 13 gradually expands with increasing temperature and pressure, the wave given by the hydrodynamic edge A retains its height and shape very well, and the sharp protective edge B the seal retains contact with the shaft and does not bulge out in any noticeable way and does not show a tendency to rise from the shaft. The measurements of the contour width were in good agreement with the predicted width of the above finite element analysis.

Analysis of the drill compaction by the finite element method showed that the geometry of the drill compaction profile, modeled at an average wave height, is relatively stable under compression and does not show a large tendency to tip over in one or the other direction. The stability of the structure is due to the fact that the geometry of the internal dynamic seal is approximately centered in the axial direction relative to another, static geometry of the seal.

Although the drill’s hydrodynamic seal is relatively stable for low pressure or non-pressure applications, it is not suitable for high pressure applications, such as the lubricated outboard hydraulic motor bearings with pressures of 500-1500 psi (3.5-10, 5 MPa), since the geometry of the dynamic seal is not directly supported by the packing box wall. In the hypothetical case, when the drill seal is exposed to high pressure grease, the seal will shift toward the opposite wall of the stuffing box. The initial contact with the wall of the stuffing box would occur at a diameter that is much larger than the sealing interface on the shaft. Hydrostatic pressure acting on a non-supported area between contact with the wall and

contact with the shaft exerted a tremendous force on the inner diameter of the seal. As a result, the inner diameter of the seal would be severely distorted by the action of a lubricant forcing it toward the opposite wall of the stuffing box. Such a strong distortion of the hydrodynamic geometry would exclude hydrodynamic lubrication of the interface and thereby lead to a rapid failure of the seal due to wear from direct friction contact, melting from frictional heating and rapid extrusion due to softening of the elastomer.

General Purpose Hydrodynamic Seal

A general purpose hydrodynamic seal has also been developed within the scope of the aforementioned US patent. 12 and 12a show a seal according to the prior art, while FIG. 12 shows the compressed state of the seal inside the packing box, and FIG. 12A shows the uncompressed state of the seal. 12 shows a typical general purpose seal E of a rotating shaft mounted in a circumferential groove of the housing, which is sized to hold an elastic circular sealing element with radial pressing against the cylindrical sealing surface of the shaft, thereby creating a static seal with the housing and a dynamic seal with a rotating shaft as well as a conventional packing type seal, such as a ring with a circular cross section. When the shaft rotates, the hydrodynamic seal remains stationary relative to the housing and maintains a static sealing interface with the specified housing, while the interface between the seal and the shaft becomes a dynamic sealing interface.

As shown in FIGS. 12 and 12a and as shown in a graph in FIG. 13, the inner peripheral surface D of the hydrodynamic seal includes a geometry that contributes to a long service life of the seal due to hydrodynamic lubrication of the dynamic zone between the seal and the shaft and by eliminating the ingress of contaminants from the seal to the shaft interface. The internal peripheral geometry of the hydrodynamic seal includes a wavy, axially variable edge A on the lubricated side of the inner diameter, and a straight or sharp protective edge B on the environment side C. With a relative rotation of the shaft, the waveform on the lubricant side L, which has a shape G gradually tapering in the axial direction, creates a hydrodynamic wedging action (due to the normal rotational velocity component Vn), as shown in Fig. 13, which leads to the lubricant film entering between seal E and shaft N. This

a lubricating film physically separates the seal and the shaft and thereby prevents the usual dry friction wear associated with conventional non-hydrodynamic stuffing box seals, which lengthens the life of the seal and the mating shaft surface and allows for higher working pressures in practice. The straight sharp edge B on the environment side C on the interface between the seal and the shaft does not change in the axial direction and does not create a hydrodynamic proppant, and therefore acts as a protection of the interface between the seal and the shaft from contamination in the form of particles. Light axial shaft movements occur in many types of rotating machinery due to the elasticity of the components and various internal clearances. An acute angle B at the edge of the non-hydrodynamic side of the sealing interface, which is commonly known as the protective side or protective edge, eliminates the scratching effect of contaminants on the shaft during such axial movements of the shaft. Thus, when relative axial movement occurs between the shaft and the seal, the accumulated contaminants, usually abrasive material, are scraped off from the shaft sealing surface, so that the dynamic sealing interface remains free from contamination. In addition, the hydrodynamic proppant, which is caused by the interaction of the lubricant and the hydrodynamic edge A, creates a controlled action of pumping the lubricant from the lubricant side to the surrounding side of the dynamic interface, which leads to a small degree of leakage of the lubricant; this leakage or movement of the lubricant film flushes particles in the form of particles from the shaft and seal surfaces to the side of the seal contamination if any contaminants penetrate the dynamic sealing surface beyond the protective edge B.

The profile shape of the rotating shaft hydrodynamic seal used in drill bits differs significantly from the profile shape that is offered for other applications. The space is very limited in the drill bit for reasons of mechanical strength, and the geometry of the profile of the hydrodynamic seal of the bit is dictated by the shape and length of the shaft interface, and the geometry and relative position of the sealing groove. These sealing constraints result in a dynamic sealing interface geometry that is approximately centered relative to the geometry of the static sealing interface. In this regard, the profile of the bit seal is significantly different from the profile of the standard Kalsi general purpose seal, which, unlike the bit seal, is designed for both high and low applications.

pressure. General purpose seals have a dynamic sealing interface that is shifted relative to the static sealing interface. Unlike the bit seal, which is designed for special applications with low pressure, the general purpose hydrodynamic seal is designed for use with both low pressure and high pressure. Therefore, the shape of the general purpose seal profile differs significantly from the shape of the bit seal profile; the sharp protective edge is preferably located at the extreme end of the seal surrounding side. The position of the sharp protective edge on the general purpose hydrodynamic seal is dictated by the need to support the packing box wall in high pressure applications. This form is caused by the fact that when conditions of high pressure of the lubricant exist, the protective side of the seal is pressed against the wall of the packing box due to the pressure difference between the lubricant and dirt, as shown in Fig. 12. If the protective edge were not at the extreme end of the seal's surroundings, the seal would experience too much distortion and the hydrodynamic characteristics would be excluded, as described previously in the hypothetical case of exposure to a high pressure seal.

Since the design of the general purpose hydrodynamic seal is such that the sharp protective edge is at the extreme end of the seal's surroundings side, and since the shape of the seal end on the surroundings side is essentially the same as the packing box wall shape on the surroundings side, the general seal has good pressure support grease in all places, except for clearance I, which exists between the housing and the shaft. This gap, often called the extrusion gap, is kept very small, so that the seal material due to its elasticity can overlap the gap and resist significant deformation due to the resulting hydrostatic pressure.

Description of the problem

Although the rotary shaft general purpose hydrodynamic seal shown in FIG. 12 works very well to contain pressure lubrication in an abrasive environment (as evidenced by its widespread use as a seal for high pressure hydraulic downhole motors), it sometimes exhibits premature abrasion when it is used to hold grease without pressure or low pressure grease in an abrasive environment.

The most common low-pressure use of a general-purpose hydrodynamic seal for a rotating shaft is

the use of hydraulic downhole motors in the sealing support nodes when drilling oil wells, which are used for drilling hard and rocky rocks, and which is a particularly difficult application of shaft compaction.

Low pressure seals in a downhole hydraulic motor must operate under the following combination of adverse conditions:

1. High levels of oscillation of lateral deflection of the shaft.

2. Increased ambient temperature due to the heat generated by geothermal factors, support and compaction.

3. Highly abrasive mud environment.

4. Limited volume of the fluid reservoir.

5. Strict restrictions on radial and axial space.

6. Static misalignment of the shaft and housing.

7. High level of vibration.

8. Axial movement of the shaft due to internal mechanical gaps and elasticity of the components.

9. Axial movement of the seal for pressure equalization reasons.

A good low-pressure seal for hydraulic downhole motors should be able to work under the combination of the above harmful conditions with low leakage and long service life. The high pressure seal for downhole hydraulic motors must also withstand the lubrication pressure of 3.5-10.5 MPa in addition to the problems mentioned above.

The hydraulic downhole motor, which is located at the lower end of the drill string, is a volumetric hydrodynamic motor that drives the drill bit. It is driven by a circulating drilling fluid, which is also used to flush cuttings out of the well.

The downhole hydraulic motor consists of three principal subassemblies: a hydrodynamic motor, a universal joint and a support unit. The circulating drilling fluid rotates the rotor of the hydrodynamic engine. The universal joint transfers rotational motion from the engine to the rotating shaft of the support subunit, to which the drill bit is connected by means of a threaded connection. The weight of the drill string is transferred to the drill bit through the thrust bearing of the support subassembly.

When the bit rotates, it abuts the geological formation and destroys it due to the weight acting on the bit, which is concentrated with the help of this

chisel cutting structure. The radial bearings of the support subunit serve to orient and direct the bit relative to the drill string.

Controlled systems use a curved housing between the engine subassembly and the reference subassembly. The direction is controlled by instantly turning the curved body in the desired direction of drilling. Vertical drilling is carried out by continuous rotation of the drill string. Curved housings add lateral load to the radial bearings of the support subunit in both direct and directional drilling modes, and together with changes in the frequency of the axial load lead to a large level of oscillation of the shaft deflection.

All housings of the sealed support assembly of the hydraulic downhole motor are filled with bearing grease, which is held by rotating sealing elements at each end of the support housing system. Lubricant pressure is balanced with drilling fluid pressure; in most designs, the pressure of the lubricant is balanced with the pressure in the borehole of the drill string. Pressure equalization is usually carried out using a pressure equalizing piston, which is made with the possibility of sliding with a seal in the housing and rotation with the seal together with the shaft. The rotary seal of the pressure equalizing piston is one of two places where general purpose hydrodynamic seals are commonly used as low pressure seals in the sealed support subassemblies of hydraulic downhole motors.

After the drilling fluid passes through the hollow shaft of the compacted support subunit, it then passes through the nozzles of the drill bit and enters the borehole ring, its pressure drops to a level that is about 3.5-10.5 MPa lower than the borehole pressure. Since the pressure of the lubricant inside the support unit of the hydraulic downhole motor is equalized with the pressure of the drill string or with the pressure of the ring, one of the rotating seals must have a pressure drop of 3.5-10.5 MPa between the lubricant and the environment. Since the lubricant pressure is usually aligned with the drill string pressure, the high pressure seal is usually located at the lower end of the assembly.

Experience has shown that it is advisable to protect the high pressure seal and the surrounding mechanical structure from the abrasive environment of the drilling fluid by providing a barrier fluid and a rotating barrier seal. Barrier Seal System, which is the second place where general purpose hydrodynamic seals are used as low seals

pressure in the sealed support units of hydraulic downhole motors, provides a clean lubricant environment on both sides of the high pressure seal.

If a barrier seal is not used, the high-pressure seal and the surrounding mechanical structure may be subject to abrasion; this wear can degrade the performance of the high pressure seal in various ways. Abrasive particles are present in the displacement gap between the housing and the shaft, which cause significant wear on the shaft and the mating hole of the housing, which leads to an increase in the displacement gap, which reduces the extrusion resistance of the rotating high-pressure seal. Abrasive wear on the shaft appears as a local groove. When relative axial movement occurs between the housing and the shaft as a result of the internal clearances of the assembly and the elasticity of the support components, the leading edge of the high pressure seal may be damaged when passing over the groove. Some drilling fluid environments may also chemically affect high pressure seal material when a barrier seal is not used.

Barrier seals are usually installed in a pressure equalizing floating piston, which is configured to sealed sliding inside the housing and sealed rotation together with the shaft. The task of the floating piston is to balance the pressure of the barrier lubricant with the ambient pressure; A barrier grease is a grease enclosed between a barrier seal and a high pressure seal.

The study of the wear of the hydrodynamic seal hydraulic downhole motors

A study of the low pressure hydrodynamic seals used in the many different configurations of the sealed support subassemblies of hydraulic downhole motors shows that premature abrasive wear of low pressure seals often occurs.

In one well-documented case, groove wear was evident on the surrounding side of the inner periphery of the low pressure seals after less than 70 hours of operation, although both seals were still effective. This wear characteristic is obviously caused by the suction of the abrasive material onto the dynamic sealing interface, i.e. abrasive particles that fall on the dynamic sealing interface through the protective edge of the seal. In another case, the low-pressure seal was abrasively worn, but still effective after more than 200 hours of operation in the hydraulic barrier seal

downhole motor. Although this characteristic looks good, many other low pressure seals from hydraulic downhole motors were noted that had completely smooth wear on the inner periphery, with no signs of leaving the original axially changing shape.

As a result of clear evidence of premature abrasion of general-purpose hydrodynamic seals when used at low pressure in hydraulic downhole motors, a test was conducted to investigate the problem under controlled laboratory conditions. The test was carried out with a general-purpose hydrodynamic seal running on the surface of finely ground tungsten carbide in the presence of a lubricant with minimal pressure. Before rotation, the seal was subjected to instantaneous lubricant pressure sufficient to orient the seal parallel to the surrounding side of the packing box. The seal surrounding side was exposed to abrasive drilling fluid containing sand. During the test, the temperature of the lubricant was maintained within the temperature range expected during operation of hydraulic downhole motors.

During the test process, the assembly was opened approximately every 24 hours to observe signs of wear. Before reassembly, the seal and shaft were thoroughly cleaned. After assembly, the seal was again subjected to instantaneous lubricant pressure to orient the seal at the surrounding side of the packing box wall.

Although the sealing surface of the interface was cleaned every day, and although the wear of the shaft was small, the seal began to show a strip of abrasive wear at about 100 hours of operation. At about 140 hours, the strip of abrasive wear increased markedly in width. At about 160 hours, the torque sharply more than doubled. The increase in torque was accompanied by a corresponding decrease in the degree of leakage. At about 170 hours, the torque fell to almost its original size, maintained this level for about 24 hours, and then began to increase. At about 200 hours, the test was stopped due to high levels of torque. At the end of the test, the seal was still effective, but showed the wear pattern shown in Fig. 14. It should be noted that on the dynamic sealing surface of the interface on the side C of the contamination of the seal, a circular wear groove J is formed.

From about 100 hours to the completion of the test on the shaft, a visible wear pattern gradually developed at the location of the interface. At about 100 hours

the shaft wear pattern was barely visible and not felt. At about 200 hours, the shaft wear pattern had a matte texture that was easily detected by touch.

It was initially thought that the matted texture of the shaft wear pattern or the suction abrasive particles were responsible for the increase in torque. This was disproved when a clean worn seal showed high torque when rotating on the surface of a new shaft. Similarly, the new seal did not show much torque when rotating on a shaft with a wear pattern. It was concluded that the increased torque that occurs at about 150 hours of testing is the result of a partial loss of hydrodynamic lubrication caused by seal wear. This conclusion is confirmed by a decrease in the degree of leakage, which occurs simultaneously with an increase in torque.

Although the seal was still effective at the end of 200 hours of testing, the barrier seals did not last for so long in actual use due to more severe conditions, such as greater dynamic wear and also lack of cleaning every 24 hours.

Along with an indication of a serious deterioration in hydrodynamic lubrication and a corresponding increase in the degree of seal wear, a sharp increase in torque in the above test has another serious consequence. Many low pressure seals in downhole hydraulic motors are installed in floating pistons. The only remedy against the rotation of the piston is the friction of the outer, sliding seal on the sealing interface between the piston and the housing. The outer, sliding seal is typically a circular cross-section ring or combination seal, such as manufactured and sold by Parker Seal under the PolyPak® trademark. Sudden increases in torque can cause the piston and sliding seal to rotate inside the housing. This rotation can be very damaging to the outer seal, since it is in dry abrasive contact with the opening of the housing. This leads to severe wear of the outer seal; Many customers have reported such wear and tear, and the inventors of this invention have observed such wear. A sharp increase in torque can also lead to the rotation of the hydrodynamic seal inside the packing box, which leads to wear of the packing and packing box due to dry sliding contact. To prevent this, many packing boxes are sandblasted to increase frictional resistance for static seal retention.

When an eccentricity condition exists between the shaft and the stuffing box, the radial compression of the seal decreases at about half the circumference of the seal and increases at the opposite half. In order to maintain seal tightness in such eccentric conditions, it is necessary to provide sufficient pre-compression so as to provide a sufficient level of compression in the shifted state. At the same time, the test maximum compression should not adversely affect the characteristics of the seal. Therefore, it is desirable that the seal can withstand pressures that vary over a wide range.

General purpose finite element analysis of hydrodynamic seals

Although finite element analysis was not used in the initial design of the existing general-purpose hydrodynamic seal, it was widely used later to investigate the exception observed in the laboratory when the shaft contact pattern indicated that part of the seal at the extreme lubricated end of the profile was slightly rubbed against the shaft during rotation . Since then, other examples of contact between the lubricated shaft end and the seal in low pressure seal samples taken from hydraulic downhole motors used in the field have been identified.

The finite element analysis was performed on a representative hydrodynamic seal and was used to evaluate the deformed shape of the seal and the resulting contact pressure between the surfaces in a wide range of conditions of temperature, compression and pressure of the lubricant. The analysis predicted an increase in distortion and torsion of the general purpose seal with increasing radial compression and a possible contact of the extreme lubricated lip of the seal with the shaft. On Fig shows a graph of the displacement of an uncompressed general purpose seal, according to the prior art, obtained using finite element analysis, while the seal was modeled at half the wave height with a compression of 11% and a temperature increase of 180 ° F (82 ° C) . The contour M shown by the dashed line represents an undeformed seal shape. The contour N shown by the solid line represents the predicted deformed shape of the seal under radial compression. The displacement graph shows that under simulation conditions the seal is twisted by such a value that the inner circular angle F of the lubricated end of the seal comes into contact with the shaft surface Q. Contact patterns of the shaft with a low pressure seal indicating this contact were observed

on samples for laboratory testing and on used low-pressure seals of hydraulic downhole motors that have worked in the field.

A finite element analysis of an incompressible seal also showed another undesirable tendency; since the seal is twisted inside the stuffing box, the sharp extruding lip B of the seal actually rises from the shaft even with a relatively low initial compression. This tendency to lift is compounded by increasing compression. On Fig shows on an enlarged scale part of the protective edge of the uncompressed General purpose seal, modeled at half the wave height with a compression of 11% and a temperature increase of 180 ° F (82 ° C), the displacement graph of which is shown in Fig. 15. On Fig shows the sharp protective edge R of the seal, raised from the opposite surface Q of the rotating shaft as a result of compression caused by twisting of the seal. As a result, the seal has a gradually tapering shape S on the surrounding side of the dynamic sealing surface T of the section. The smoothly tapering shape of the seal is reflected in the contact pressure profile between the surfaces predicted by finite element analysis.

In the ideal case, the contact pressure profile on the surrounding side of the dynamic sealing interface should rise sharply at the sharp edge of the seal, since the sharp edge is designed to provide a protective scraping action. There is no scraping effect when a general purpose hydrodynamic seal is used in a pressure-free application; instead, the local seal geometry takes the form of the aforementioned gradual approach to the shaft, which is ideal for trapping abrasive particles in the event of axial movement of the shaft. A small axial movement of the shaft is present in the sealed support nodes of the downhole hydraulic motors and in other types of equipment due to clearances in bearings and other mechanical clearances and elasticity of the components. As a result of the rise of the sharp edge of the seal and the resulting smoothly tapering local seal geometry, the axial movement leads to the movement of the abrasive particles on the dynamic sealing interface. This causes abrasive wear on the shaft and seal.

The degree of torsion of the seal inside the packing box depends on the localized height of the hydrodynamic wave. The tipping and corresponding rise of the sharp edge is greater at the bottom of the wave and less at the top of the wave. This means that the actual last contact point on the surrounding side of the dynamic sealing interface changes its position in the axial

direction depending on the height of the hydrodynamic input wave. This phenomenon, unfortunately, creates a limited hydrodynamic effect of the protective edge, which acts to wedge the surrounding dirt into the sealing surface of the interface and causes premature seal wear even in the absence of axial movement of the shaft, and thereby limits the life of the seal. This conclusion is confirmed by worn out seals from laboratory tests, where there was no axial shaft movement.

The above finite element analysis model with the same compression of 11% and a temperature increase of 180 ° F. (82 ° C.) was used at a pressure difference of 100 psi (690 kPa), as shown in FIG. 16. This load case predicts that the pressure on the lubricant side causes the seal to straighten and abut against the wall surroundings. This brings the sharp edge into the desired contact with the shaft, which leads to the desired sharp contact pressure profile on the surrounding side of the sealing interface and the desired protective scraping action. The rectifying effect of lubrication pressure also has the desired effect of eliminating shaft contact at the lubricated end of the seal, although the predicted clearance is still small.

A general purpose seal was also modeled with a compression of 15% and a temperature difference of 180 ° F. (82 ° C.), as shown in FIG. At this compression level, as shown by the solid line, indeed the entire inner part of the seal became flat in contact with the shaft; the seal model remained really completely flat at the shaft even when a pressure difference of 690 kPa was applied to the grease side. This extreme distortion of the hydrodynamic geometry strictly prohibits or completely eliminates the desired effect of the hydrodynamic pumping of the seal, since the hydrodynamic lip G is really isolated from the lubricant due to the contact of the circular lip F with the shaft.

Based on the results of the finite element analysis, it was concluded that the initial compression of a general purpose seal should be limited to less than 9% in the absence of pressure, although compression of 12% should be used if a lubricant pressure of 690 kPa or more is present. Based on finite element analysis, Kalsi Engineering, Inc. started recommending a nominal initial compression of 7.5%. It is clear that it is desirable to have seals that can withstand higher levels of compression, since such a seal would allow misalignment, runout, tolerances, permanent compression deformation and wear.

As a general rule, seals with a smaller profile must be subjected to greater initial compression than larger seals, since any given degree of misalignment, dynamic runout, tolerances, wear, etc. corresponds to a higher percentage of a smaller seal profile. The compression limitations of the existing seal design impose appropriate restrictions on the minimum effective seal profile size, which is practical to manufacture.

In general, finite element analysis predicts that the torsional effect of uncompressed seals that occurs when a radial pressure is applied to the seal eliminates the desired scraping effect of the protective sharp edge and causes an undesirable hydrodynamic effect on the interface between the seal and the shaft on the contaminated control side, which exhibits a tendency for injection of abrasive particles onto the sealing interface. The resulting abrasive wear of the seal due to the presence of abrasive particles and due to the minimized flushing action of the lubricant, which is otherwise observed, limits the life of the seal.

There is no tendency to raise the protective lip in pressure applications, since the lubricant pressure presses the seal against the opposite wall of the stuffing box, which changes the orientation of the seal and presses the protective edge against the shaft. Although a compressible seal is not sensitive to the torsional effect observed in uncompressed seals, both seals experience extremely high distortion at high compression levels and exhibit a concomitant exclusion of hydrodynamic action. It would be desirable to have higher compression levels than those currently achieved.

Field tests of a conventional hydrodynamic seal operating under conditions of low fluid pressure and high mechanical compression confirm excessive abrasive wear after a relatively short operating time: therefore, it can be assumed that such seals are distorted and twisted during mechanical compression so that the sharp protective edge rises from the sealing the surface of the shaft, as shown in Fig. 18, which allows abrasive particles to penetrate under the raised sharp edge.

The final service life of the available hydrodynamic shaft seals is also limited by the residual compressive deformation of the elastomer (permanent deformation) during installation with high temperature. It has been found that most of the permanent compression deformation usually occurs in a protruding dynamic

a sealing lip on the inner periphery of the seal, and not across the main rectangular seal housing. The reason for this phenomenon is twofold. Firstly, the protrusion is located directly at the source of self-generated heating, i.e. shear lubricant in the sealing surface section. Secondly, since the protrusion is much smaller than the main body of the seal, it shrinks in percentage terms more than the main body due to the difference in stiffness. Resistance to permanent compression deformation of a seal can be improved if the distribution is distributed more evenly over the seal body instead of being concentrated in the protruding protrusion.

SUMMARY OF THE INVENTION

The invention relates to a hydrodynamic seal for a rotary shaft seal, which eliminates the problems listed above with respect to a general purpose hydrodynamic seal for a rotary shaft, designed to operate with lubrication under pressure and without pressure in severe abrasive conditions. The hydrodynamic seal of a rotating shaft according to this invention is provided with an axially varying shape on the lubricant side of the dynamic sealing interface for the purpose of hydrodynamic lubrication of the dynamic sealing interface and has a straight, acute-angled, axially non-changing protective shape on the surrounding side of the dynamic sealing interface in order to eliminate pollution. In particular, the hydrodynamic seal of the invention relates to a commercially available type of rotational shaft hydrodynamic seal that is manufactured and sold by Kalsi Engineering, Inc. of Sugarland, Tx. in accordance with US patent No. 4610319. In this hydrodynamic seal, a protective lip is located at the extreme end of the profile geometry in order to provide mechanical support against the lubricant pressure acting on the hydrostatic area of the seal to create a profile geometry that is compatible with applications of the high-pressure grease rotating seal.

The present invention dramatically improves the already exceptional characteristics of said general purpose hydrodynamic rotary shaft seals, particularly when seals are used in non-pressure lubricating or low pressure lubricating applications. Improvement is achieved by eliminating the currently observed tendency to raise the protrusion.

Improving the characteristics of the security feature leads to a significant increase in the service life when using seals in abrasive conditions. Improving the protective characteristics also makes the seal less sensitive to strong fluctuations in torque, which leads to rotation of the seal inside the packing box; when the seal is installed in the floating piston system, the likelihood that the improved seal causes the piston to rotate in its bore is reduced.

The invention also extends the life of seals by providing higher levels of initial compression without compromising the sign of hydrodynamic lubrication of the seal; this additional compression makes the seal significantly less sensitive to permanent compression deformation, abrasion, mechanical misalignment, dynamic runout and manufacturing tolerances.

The invention also reduces contact pressure between surfaces near the input geometry of the hydrodynamic lubricant, which makes the hydrodynamic lubricant more complete and efficient at low speeds.

The invention also provides moderate additional improvements in resistance to permanent compression deformation of the seal due to the redistribution of compression over a large part of the seal.

The invention provides all of the above advantages due to the implementation of a simple and compact protruding sealing geometry at a non-dynamic sealing interface, which provides resistance to the compression load without causing the corresponding twisting of the seal profile. The resulting improved seal can be made using existing molds with only minor modifications to the tool. The improved seal requires less raw materials for manufacturing, and is included in the same packing box as the existing general-purpose hydrodynamic seal, and therefore does not require a change in existing equipment during replacement.

The invention relates to a hydrodynamic seal for a rotating shaft having a convex geometry of a static sealing interface and adjacent geometry of the seal profile, which essentially repeats and is a mirror geometry of a dynamic sealing surface. With this arrangement, the shape of the overall profile of the hydrodynamic compaction approaches the condition of two-sided symmetry, which has a much less tendency to twist during compression, since the distortion of the profile caused by compression on

a static sealing interface, approximately the same as the distortion caused by compression on the dynamic sealing interface. This symmetrical seal configuration does not tend to cause the seal to twist in the packing box when the seal is radially compressed.

A standard rotary shaft general purpose hydrodynamic seal, as well as a rotary shaft hydrodynamic seal, incorporating the convex geometry of a static sealing interface, were tested in a laboratory using a lubricant holding pressure of 15 psi (100 kPa) under abrasive drilling fluid. The seals were tested on separate parts of the same shaft surface, with the same runout, the same stuffing box and the same temperature and speed. Tests of a standard general purpose seal were described in detail in the Problem Description section of this description. In short, the general purpose seal was removed for inspection every 24 hours, cleaned and reinstalled. It showed abrasive wear in the form of a groove at 110 hours and a sharp increase in torque and an accompanying decrease in the degree of lubricant leakage at about 160 hours of operation. In contrast, a seal that includes a convex static sealing lip according to this invention was disassembled for the first time after 144 hours, while it did not have the advantage of daily cleaning, and yet showed absolutely no wear. The test was completed after 508 hours of operation, and the seal and shaft were thoroughly checked. Neither the seal nor the shaft showed any signs of wear. Torque remained small and constant throughout the test. The test confirmed that the hydrodynamic seal of the rotating shaft according to this invention provides truly extraordinary performance improvements over the standard general purpose hydrodynamic seal.

For a better explanation and understanding of the above features, advantages and objectives of the present invention, a detailed description of the invention is given below with reference to the embodiments shown in the accompanying drawings.

However, it should be noted that the accompanying drawings only illustrate typical embodiments of the present invention and therefore should not be construed as limiting the scope of the invention, since the invention can be applied to other also effective embodiments.

Brief Description of the Drawings

figure 1 is a partial section of the housing system and the rotating shaft, including a hydrodynamic seal of the rotating shaft, made in accordance with this invention;

FIGS. 2 and 2A are a partial cross-sectional view of a hydrodynamic seal of a rotating shaft according to a preferred embodiment of the present invention, wherein FIG. 2A shows the compressed state of the seal inside the stuffing box, and FIG. 2A shows the non-compressed state of the seal;

figure 3 is a section of a hydrodynamic seal of a rotating shaft, according to an alternative embodiment of the present invention;

4 and 4A are a partial sectional view of a housing system and a rotary shaft including a hydrodynamic seal of a rotary shaft according to the present invention, wherein FIG. 4 shows a seal inside the inner groove of the housing for a seal, and FIG. 4A shows an uncompressed seal configuration;

5 is a sectional view of a hydrodynamic seal of a rotating shaft, according to another alternative embodiment of the present invention;

6 and 6A are a partial sectional view of a hydrodynamic seal according to an alternative embodiment of the present invention, wherein FIG. 6 shows a seal radially compressed inside a seal groove having an inclined outer abutment surface, and FIG. 4A shows an uncompressed seal configuration;

FIGS. 7 and 7A are a partial cross-sectional view of a hydrodynamic seal according to an alternative embodiment of the present invention, wherein FIG. 7 shows the seal in its compressed state and FIG. 7A shows an uncompressed seal;

FIGS. 8 and 8A are a partial cross-sectional view of a hydrodynamic seal according to another embodiment of the present invention, wherein FIG. 8 shows the seal in its compressed state, and FIG. 8A shows the seal in its non-compressed state;

FIGS. 9 and 9A are a partial sectional view of a hydrodynamic seal according to another alternative embodiment of the present invention, wherein FIG. 8 shows the seal in its compressed state, and FIG. 8A shows an uncompressed configuration of the seal;

figure 10 and 10A is a partial section of a hydrodynamic seal, according to another variant implementation of the present invention, while figure 10 seal

shown in its compressed state, and in FIG. 10A, the seal is shown in an uncompressed state;

11 is a partial section of the hydrodynamic seal of the drill bit, according to the prior art;

12 and 12A are a partial sectional view of a general purpose hydrodynamic seal, according to the prior art, wherein in FIG. 12 the seal is in a compressed state inside the stuffing box to form a hydrodynamic sealing system with a rotating shaft, and FIG. 12A shows a seal in an uncompressed condition;

Fig - configuration of the sealing surface of the interface of the hydrodynamic seal, which illustrates the theoretical course of development of the action of wedging or forcing;

Fig - hydrodynamic seal of a rotating shaft of General purpose, according to the prior art, with a pattern of wear caused by the absorption of abrasive material from the side of the seal contamination;

15, 16, 17 are graphs of finite element analysis of general-purpose hydrodynamic seals, wherein the uncompressed state of the seal is shown by a dashed line, and the radial compression state is shown by a solid line;

Fig. 18 is an elevation of the seal lip from a rotating shaft by tipping or twisting the seal in response to radial compression at a low pressure difference, on an enlarged scale.

Detailed Description of a Preferred Embodiment

Figures 1, 2 and 2A show a section through a hydrodynamic seal assembly of a rotating shaft, indicated generally by 10, including a housing 12 from which the rotating shaft 14 exits. The housing defines an internal groove for sealing, a saddle or stuffing box 16, inside of which there is a hydrodynamic a rotating shaft seal 18, which is made in accordance with the principles of the present invention and which is shown in more detail in partial section in FIGS. 2 and 2A. The configuration shown in FIG. 2A is a sectional view of an improved hydrodynamic seal of a rotating shaft according to this invention in a radially uncompressed state, while FIG. 2 shows a sectional configuration of a seal located inside a groove and radially compressed between the rotating shaft and the radially outer wall grooves for sealing. The radial section shown in figure 2 and 2A, made in place

the periphery, which represents the average width of the contact between the surfaces of the dynamic sealing interface, which corresponds to the midpoint of the wave height of the geometry, which creates the effect of wedging in the lubricant described in US patent No. 4610319.

In terms of general orientation, the end of the seal, which is oriented in the direction of lubrication, is surface 20. This surface may differ from the flat surface shown in FIGS. 2 and 2A, without departing from the idea and scope of the present invention. The end of the seal, which is oriented in the direction of the hard abrasive or contaminated environment 22, is surface 24. This surface may also differ from the flat surface shown in FIGS. 2 and 2A, without departing from the idea and scope of the present invention.

According to the present invention, a circular hydrodynamic sealing element 18 is installed inside the circular seat or groove 16 for sealing, and is pressed against the rotating surface 28 of the rotating shaft 14. When the seal is installed in the circular groove or seat for sealing, the circular radially protruding static sealing protrusion or bulge 30 pressed against the opposite surface 32. On the inner periphery of the circular sealing element 18, an inner circumferential bulge 34 is provided, which there is a dynamic sealing surface 36, which is pressed against the opposite rotating surface 28, which usually, but not always, is a circular sealing surface of the rotating shaft, such as the surface indicated by the position 14. Hydrodynamic sealing element 18 is supported with sufficient radial compression to ensure tight against liquid seals on the static sealing interface between the static sealing protrusion 30 and the mating opposite The surface 32 of the seat, and between the dynamic sealing lip 34 and the opposite surface 28 rotating relative thereto. A hydrodynamic seal is used to separate the lubricant in the lubrication chamber or the lubrication side 38 of the seal from the contaminated fluid of the environment 22 and to prevent mixing of the lubricant and the contaminated material inside the environment. In the field of well drilling, the contaminated fluid is usually a drilling fluid called a “drilling fluid” that contains highly abrasive particles within the fluid carrier.

The annular hydrodynamic sealing element may consist of any number of suitable sealing materials, including elastomeric or

rubber-like sealing material and various polymer sealing materials.

The inner and outer radial bulges form conical surfaces 40 and 42, respectively, which are directed to the lubrication side of the seal, while the surface 40 has a non-circular configuration, while the surface 42 has a circular configuration. The conical surfaces 40 and 42 define each radius, which determines the gradual convergence with the shaft and seal, respectively. The conical surface 40 and the inner cylindrical dynamic sealing surface 36 are mated with a radius 44. The surface 40 intersects the cylindrical surface 46 at an acute intersection angle 48. The conical surface 40 and the radius 44 form an angular, gradually tapering shape on the lubricant side of the dynamic sealing interface, which causes the hydrodynamic migration of the lubricant film to the dynamic sealing interface, as described in US patent No. 4610319. The geometry of the conical surface 40 may take any of many suitable forms, which provides a gradually tapering shape on the lubricant side of the dynamic sealing lip or bulge, without departing from the idea and scope of the present invention. A gradually tapering shape changes position along the periphery of the seal so that the lubricant film is hydrodynamically wedged into the dynamic sealing surface with a relative pivotal movement between the lubricant and the seal.

The circular angle or protective edge 50 represents the geometry of the steep scraping edge of the hydrodynamic seal, which does not develop a hydrodynamic action with the environment in response to the relative rotational movement and which scrapes the dirt from the rotary shaft during the relative rotational movement between the shaft and the seal. Thus, the seal eliminates the entry of abrasive particles present on the contamination side of the seal onto the dynamic sealing interface.

Figures 2 and 2A and the invention shown therein relate, in particular, to the conventional and customary type of general purpose hydrodynamic seal of a rotating shaft, in which the location and configuration of the edge 50 and the end 24 of the surroundings of the seal 18 are such that they are strongly supported by the wall 52 of the packing box so that it provides resistance to the ejection of the seal material and other related damage to the seal in cases where the seal is affected by hydrostatic force caused by the pressure of the lubricant, acting on

the area between the static sealing interface and the dynamic sealing interface.

The combination of the sharp protective angle 50 and the angular gradually tapering geometry of the conical surface 40 and the radius 44 interact with the formation of a convex protrusion 34, which protrudes radially inward from the main body of the seal 18 at the extreme end of the seal, which represents the surrounding side 24 of the seal profile. The convex protrusion 34 during compression forms a dynamic sealing interface in cooperation with a rotating opposite surface 28 formed by the shaft 14.

A principal feature of the geometry of a preferred embodiment of the present invention is a convex static sealing lip 30, which protrudes radially outward from the main body of the seal 18 at the extreme end 24 of the seal, which represents the surrounding side of the seal profile. The convex static sealing lip 30 when compressed creates a static sealing interface in cooperation with the opposite surface 32 of the circular landing groove.

In the preferred embodiment shown in FIGS. 2 and 2A, the convex static sealing lip 30 according to this invention has a substantially complete profile configuration identical to the average profile configuration of the convex dynamic sealing lip 34, except that the overall protrusion 54 of the static sealing the protrusion 30 is slightly smaller than the overall protrusion 56 of the dynamic sealing protrusion (the width of the convex static sealing protrusion 30 can also be changed to fit the changing width of the dynamic protrusion at any given location on the periphery). Ideally, the protrusion 54 of the static sealing lip should be equal to or slightly less than half the nominal radial compression of the seal, and the protrusion 56 of the dynamic sealing lip should be slightly larger than half the nominal radial compression of the seal.

Since the protrusion 54 of the static sealing lip is less than or equal to half the nominal radial compression of the seal, as shown in FIG. 2, the majority or the entire surface of the seal from the circular angles 58 to 60 is brought into direct contact with the adjacent opposite surface 32 of the landing groove when the seal is compressed. The close proximity and / or contact between the seal and the seating groove near the circular angle 58 provides resistance against clockwise twisting of the seal inside

stuffing box, while the clockwise direction corresponds to figure 2. This feature of stabilization of the seal is especially important in applications where hydrostatic and / or mechanical force is applied to the end 24 of the seal environment, for example, during transient pressure fluctuations or when the protective edge 50 of the dynamic sealing interface is required for active scraping of the axially moving contaminants shaft direction.

Since the protrusion 56 of the dynamic sealing lip 34 is larger, preferably substantially larger than half of the nominal radial compression of the seal, the seal surface at the radially inner circular angle 62 does not come into contact with the opposite surface 28 rotating relative to it, the angular hydrodynamic input geometry from point 48 to point 44 is not completely flat on a rotating opposite surface and the desired hydrodynamic wedging of the lubricant into the dynamic seal The real interface is not delayed by any undesirable distortion by the hydrodynamic geometry. This general trend means that hydrodynamic seals using a convex static sealing lip according to this invention can be subjected to higher levels of initial compression compared to hydrodynamic seals of a conventional design. Higher levels of compression allow using the present invention to make improved hydrodynamic seals less sensitive to mechanical misalignment, dynamic runout, permanent compression deformation and mechanical tolerances, and makes it more practical to make seals having smaller profiles.

In a preferred embodiment, according to FIGS. 2 and 2A, the overall profile shape of the convex static sealing lip 30 is almost identical to the general profile shape of the convex dynamic sealing lip 34; therefore, when compressing the seal, the profiles of the contact forces between the surfaces and the deformation of the two protrusions are very similar both in magnitude and location, and as a result, there is generally no tendency to twist the seal inside the packing box to achieve an equilibrium state.

Since the tendency to twist inside the stuffing box is really eliminated by the present invention, any tendency to raise the sharp edge 50 from the opposing rotating surface 28 is accordingly eliminated. This means that the sharp sharp edge 50 remains firmly pressed against the shaft and can fulfill its intended scraping function and protection against

penetration of contaminants from the environment onto the dynamic sealing interface. Improved protective effect means that the penetration of the abrasive material onto the sealing interface is significantly reduced and, accordingly, the abrasive wear of the seal and the associated rotating opposite surface is minimized. Therefore, as the tests confirm, the service life of the seal is significantly increased, and the torque and the degree of hydrodynamic leakage remain very uniform during the effective service life of the seal due to the absence of wear. The relative absence of shaft wear means that there is no need for repeated required coating and grinding of the shaft, which leads to significant cost savings. Since the torque is more stable with the seal according to this invention, the cases of rotation of the seals inside the packing box due to friction between the seal and the shaft are minimized; and as a result, there is no longer any need for a special improvement in the packing box friction, such as the sandblasting currently being performed.

Due to the significant elimination of the tendency for the seal to twist inside the stuffing box, the dynamic contact pressure between the surfaces near the gradually narrowing hydrodynamic input geometry is reduced, since the seal material at this point no longer counteracts the twisting action of the entire profile. Reducing the contact pressure at this critical point makes the hydrodynamic lubricant more complete and effective at lower speeds.

Due to the presence of two radially opposite convex protrusions, one static and the other dynamic, the actual compression of the dynamic protrusion is reduced in comparison with the compression of the dynamic protrusion in the existing general-purpose hydrodynamic seal. This provides improved resistance to permanent compression deformation, since most of the residual compression deformation currently occurs in the dynamic protrusion due to its proximity to the heat generated on the dynamic sealing interface.

Hydraulic downhole motors with sealing bearings are increasingly used in “hot” wells, in which the temperature sometimes exceeds the operating temperature limits of elastomers commonly used in hydrodynamic seals. The elastomers currently used for hydrodynamic seals are selected primarily on the basis of high abrasion resistance and usually have a maximum operating temperature in the range from 250 ° F to 320 ° F

(120-160 ° C). In the hydrodynamic seals according to this invention, elastomers that have higher temperature limits but less abrasion resistance, such as fluoroelastomers, can be used, since the torsion-resistant seal protects the sealing interface much better from abrasive materials.

The present invention is applicable where the rotating shaft is sealed relative to the housing, wherein the housing or shaft is a rotating member. Convex protrusions may be on the inner or outer side of the seal profile when the seal is compressed in the radial direction, with the dynamic protrusion located on the inner or outer periphery. Alternatively, the convex protrusions may be at opposite ends of the profile shape when the seal is axially pressed against flat opposing surfaces rotating relative to it.

Although the dynamic and static bulges protrude all the way to the end of the contamination of the seal, as shown in the figures, this is not necessary within the idea and scope of the present invention. It is only necessary that these bulges have such a relative position that the radial compression of the seal under low pressure conditions does not lead to twisting of the seal inside the stuffing box.

Description of Alternative Embodiments

FIG. 3 shows an alternative embodiment in which the hydrodynamic seal 66 shown is substantially identical to the embodiment shown in FIG. 2A in all respects, except that the sharp circular angle 68 replaces the rounded corner 64 of the seal shown in figa. The acute angular angle provides a protection function to help prevent grease from being injected onto the static sealing interface when the seal moves back and forth inside the packing box due to fluctuations and / or reversal of pressure conditions, which further reduces the tendency for the seal to rotate. inside the stuffing box.

It is understood that the gland type elastic sealing element is in itself an incomplete sealing device and cannot fulfill the desired sealing function without interacting with the corresponding packing box functions. Therefore, a complete hydrodynamic sealing device consists of a combination of a stuffing box and an elastic sealing element. 4 and 4A show, respectively, a compressed and uncompressed state of an alternative embodiment

inventions. As shown in FIG. 4, the sealing geometry of the packing box 70, which, in conjunction with the convex static sealing protrusion 75 of the hydrodynamic packing member 76, helps to prevent the packing from being installed in the opposite direction and the packing moving back and forth inside the packing box. In this system, the radial protrusion 72 of the static sealing lip is significantly increased compared to the seal shown in FIGS. 2, 2A and 3, which results in a protrusion that is more than half of the initial compression. A corresponding groove 74 is formed in the landing stuffing box to accommodate the protrusion 75 of the static sealing protrusion and compress it in the radial direction to a value that is approximately half the total radial compression. In the compressed state, the tail end 77 of the seal maintains sealing contact with the seat in portion 78, so that it helps prevent clockwise twisting of the seal due to external loading, while a circular angle 78 provides anti-tipping interaction with the cylindrical bearing surface 79. If the seal shown figure 4, is accidentally installed in the opposite direction, if it can be installed at all so, it will protrude from the stuffing box so far that the installation of the shaft will be extremely difficult on or physically impossible, thereby drawing attention to the fact that the seal is accidentally installed backwards. In addition, both the geometry of the static sealing lip and the corresponding geometry of the stuffing box provide a visual hint for proper orientation before installation. After installation, the convex static sealing lip 75 is gripped by the corresponding geometry 74 formed in the sealing groove. The associated geometry of the seal 76 and seal groove 70 also limits axial movement there and back in response to pressure reversal. In addition, this conjugate seal and groove geometry increases the structural integrity of the seal and thereby allows an increase in the radial compression of the seal compared to other hydrodynamic seals.

FIG. 5 shows another partial cross-sectional view of a profile configuration of a hydrodynamic sealing element 80, representing another alternative embodiment of the present invention. The seal configuration shown in FIG. 5 represents an easy modification of the seal configurations shown in FIGS. 2, 2A and 3, wherein the convex static sealing lip 82 is configured as a conical portion of the seal that forms a chamfer. Although not as effective as the profile shown in FIGS. 2, 2A and 3, due to the greater difference

between the geometry of the static sealing protrusion 82 and the dynamic sealing protrusion 84, shown in FIG. 5, the shape still provides the same general advantages as the seal and the embodiments shown in FIGS. 2, 2A and 3, although, perhaps, in lesser extent.

As shown in FIGS. 6-10, controlled radial compression of the hydrodynamic seal of the rotating shaft can be accomplished by interacting with the configuration of the inner surface of the packing box and the outer configuration of the hydrodynamic seal. In each of the embodiments shown in FIGS. 6-10, the radial compression of the hydrodynamic sealing element is ensured by its substantially symmetrical deformation on both the static and dynamic sealing surfaces, so that the radial compression does not cause the sealing element to twist inside the stuffing box.

Figures 6 and 6A respectively show a radially compressed and uncompressed seal configuration, wherein a housing 90 is shown for defining a packing box 92 having a tapered or truncated tapered radial surface 94. The hydrodynamic sealing element also forms a tapered radially outer surface 98 having stiffer taper compared to surface 94. Seal 96 is included in gland box 92 in an uncompressed state, so that radially outer corner 100 of seal 96 is in close contact with the inner m, the radially outer corner 102 of the packing box. Thus, when the rotating shaft 104 is inserted through the seal, the seal deforms both at its hydrodynamic bulge 106 and at the part that is radially opposite to the hydrodynamic bulge, so that a substantially symmetric radial compression of the seal occurs. When the seal is compressed in this way, it has no tendency to tip over or twist inside the stuffing box, and thus the protective lip 108 remains in optimum contact with the cylindrical surface of the shaft 104, which protects it from dirt, in addition to that shown in FIG. 6 in the compressed state, the outer circumferential angle 99 of the seal is in contact with the abutment surface 98 of the circumferential groove for the seal and thereby obtains structural support against compression-induced tipping inside the groove for the seal. Due to the taper of the surface 94, the radial compression of the seal attempts to squeeze the seal in the direction of the inner circular angle 102 of the seal groove, thereby causing the contamination side of the seal to be pressed firmly against the supporting surface 103.

The configuration of the interaction of the seal 96 with the seal 92 also prevents the installation of the seal in the opposite direction inside the packing box. When installed in the opposite direction, when the hydrodynamic surfaces are directed to the side of contamination, the seal is excessively compressed in the radial direction so that it becomes difficult or impossible to install the shaft. The hydrodynamic seal according to FIGS. 4 and 4A is also difficult or impossible to install in the opposite direction for the same reason.

As shown in FIGS. 7 and 7A, the housing 130 is provided with an inner stuffing box 132, which defines an inner compression surface 134 that creates an inner relief region 136 of the stuffing box relative to a portion 138 of the outer surface of the hydrodynamic seal 140 of the rotary shaft. To facilitate understanding of an embodiment, the seal 140 is shown in its radially compressed state in FIG. 7 and in an uncompressed state in FIG. 7A. On figa shows the configuration of a hydrodynamic seal General purpose, which is deformed into another configuration under the influence of radial compression. The seal compression surface 134 is radially opposite the hydrodynamic internal convexity of the seal 142 and has a width substantially equal to the compressed average width of the dynamic hydrodynamic convex seal surface 144.

Thus, after radially compressing the seal 140 inside the packing box 132 with the shaft 146 installed, the radially outer static portion of the seal and the radially inner hydrodynamic bulge 142 undergo substantially symmetrical compression and deformation. This symmetric radial compression ensures that the hydrodynamic sealing element 142 does not tip and twist as a result of radial compression. Thus, the sealing element can be subjected to significantly greater radial compression than was acceptable for commercially available hydrodynamic sealing elements. The increased compression capability provided by the combined configurations of the hydrodynamic sealing element and the packing box ensures that the protective edge 142 of the sealing element is kept in contact with the cylindrical sealing surface of the shaft 146, which prevents penetration of particles. In addition, the radial deformation of the seal is such that the radially outer angle 139 the seal is located in contact with the bearing surface 141 of the stuffing box to provide support a tilting seal within the stuffing box.

As shown in FIG. 8, the structure of the housing 150 defines an inner stuffing box 152 that forms a tapered or truncated conical surface 154 of the stuffing box having a smallest diameter 156 corresponding to the outer diameter 158 of the hydrodynamic sealing member 160. When exposed to a radial compression member, as shown in figa, the outer peripheral part of the sealing element is deformed by pressing against the conical surface 154 of the stuffing box, and also deforms hydro dynamic internal bulge 162. Since the sealing element deforms only on the dirt side to form a static seal with surface 154, the compression deformation of the sealing element is essentially symmetrical and the sealing element does not tend to tip or twist inside the packing box 152. If there is no twisting, the protective the lip 164 maintains its optimum adhesion to the cylindrical outer sealing surface 166 of the rotatable shaft. Additionally, the outer seal angle 151 is supported by the abutment surface 154, as shown in FIG.

As shown in FIGS. 9 and 9A, the hydrodynamic sealing element 168 is substantially in the same configuration as the general purpose hydrodynamic seal of a rotating shaft, such as that shown in FIG. 8, as shown in FIG. 9, in a radially compressed state inside the inner groove 170 for sealing the housing 172. The housing structure forming the groove for sealing defines a substantially cylindrical compression compression part 174 that extends laterally a distance substantially equal to the average compressed width of the hydrodi the sealing surface 176. The stuffing box includes a tapered inner surface 178 that interacts with the outer peripheral surface 180 of the sealing member 168 to define a circumferential relief area. When a radial compression is applied to the seal between the cylindrical sealing surface 184 of the rotatable shaft and the radially opposite compression surface 174, the seal deforms substantially symmetrically and does not tend to tip or twist inside the packing box 170. The contaminant lip 186 remains in scraping contact with the cylindrical surface shaft and provides effective protection against the penetration of contaminants onto a dynamic sealing interface . The radially outer seal angle 181 is supported by the abutment surface 178,

thereby supporting the seal against twisting clockwise inside the stuffing box.

Under the influence of radial compression, the hydrodynamic sealing element can be deformed so that its dynamic outer periphery assumes the optimal configuration for effective wedging of the lubricant and eliminating the penetration of contaminants. 10 and 10A show the housing structure 188 through which the rotating shaft 190 passes. The housing defines an inner packing box 192 inside which an annular hydrodynamic sealing member 194 is located. An uncompressed configuration of the sealing member 194 is shown in FIG. 10A, while how the compressed configuration of the seal is shown in FIG. 10. In the state of the hydrodynamic seal shown in FIG. 10A, the inner surface 195 and the hydrodynamic seal surface 196 of the uncompressed seal are angled relative to the horizontal. The housing structure forming the packing box 192 defines a generally cylindrical radial outer surface 198 that biases the radially outer part of the sealing element, as shown in FIG. 10. At the same time, this radial compression biases the sealing element so that its inner periphery adopts the configuration of the rotating shaft, as shown in FIG. 10, which leads to the adoption by the inner periphery of the seal of the desired configuration for effective hydrodynamic wedging of the lubricant and allows the circular lip 197 to be at a distance from the shaft, as shown in FIG. 10. In this case, it should be taken into account that the final compressed configuration of the seal is somewhat twisted, however, this twisted configuration is optimal for the hydrodynamic movement of lubricant on a dynamic surface and for effective compression in order to effectively seal on a dynamic sealing surface and for static sealing on the radially outer part of the sealing element. In addition, the protective edge 200 of the sealing element remains in effective scraping contact with the cylindrical surface of the shaft 190. Thus, although the sealing element is twisted by radial compression, the internal configuration of the sealing element compensates for this to keep the sharp protective edge in contact with the contaminant shaft sealing surface.

In view of the foregoing, it is understood that the invention is well suited to achieve the above objectives and features, together with other goals and features inherent in the device disclosed herein.

As is understood by those skilled in the art, this invention can be carried out in other special forms without departing from the idea or essential characteristics. Therefore, this embodiment should be considered as having an illustrative and not restrictive character, while the scope of the invention is determined by the claims and not by the previous description, and therefore all changes within the meaning and limit of the equivalent of the claims should be covered by it.

Claims (9)

1. Torsion-resistant hydrodynamically lubricated rotating packing element of a stuffing box type for a lubricated seal-free seal with a circular sealing surface rotatable relative to it, comprising: (a) a substantially annular body of elastic sealing material having inner and outer radial peripheries, one of which is configured to create a dynamic sealing interface with a circular sealing surface rotatable relative to it, and the other of which is configured to create a peripheral sealing contact with a static sealing surface of the packing box located at a radial distance from the indicated rotatable circular sealing overhnosti, said generally annular body further having axial end portions defining a lubricant side and a contaminant side; (b) a dynamic circumferential sealing bulge provided on one of said inner or outer radial peripheries of said annular body and configured to seal interaction with said rotatable circular sealing surface, said dynamic circumferential sealing bulge having an axially variable configuration on said side lubricants to create a hydrodynamic wedging of the lubricant inside the specified dynamic sealing o the interface and for setting a sharp circular edge on the specified side of the contaminants to prevent the penetration of contaminants on the specified dynamic sealing interface; (c) a circular static circumferential sealing bulge provided on another of said inner and outer radial peripheries of said annular body of elastic sealing material and arranged for sealing contact with said static sealing surface of the packing box, wherein said circular static circumferential sealing bulge is radially opposite the specified dynamic circumferential sealing bulge; and (d) wherein said dynamic and static circumferential sealing protuberances are substantially symmetrical and suffer substantially the same distortion during radial compression of said substantially annular body of elastic sealing material between said rotatable circular sealing surface and said static sealing surface of the packing box, due to whereby the specified mainly annular body of elastic sealing material remains free from the resulting compression Thieme twisting when exposed to radial compression. 2. The torsion-resistant hydrodynamically lubricated rotary stuffing box-type sealing element according to claim 1, wherein said static circumferential sealing convexity and said dynamic circumferential sealing convex interact to provide resistance to compressive loads without causing a corresponding twisting of the generally ring-shaped body from the sealing material forming the specified hydrodynamically lubricated rotating sealing element. 3. The torsion-resistant hydrodynamically lubricated rotating stuffing box-type sealing element according to claim 1, wherein said static and dynamic circumferential sealing bulges are both located on the indicated side of the impurities of said generally annular body of elastic sealing material and arranged radially opposite. 4. The torsion-resistant hydrodynamically lubricated rotating stuffing box-type sealing element according to claim 3, wherein said static and dynamic circumferential sealing bulges define each sharp circular protective edge on the indicated side of the contaminants of said generally annular body of sealing material, wherein said circular protective edges are located in a common plane. 5. The torsion-resistant hydrodynamically lubricated rotating stuffing box-type sealing element according to claim 1, wherein a housing is provided that defines a circular packing box that forms an internal surface configuration for compressive sealing contact with said static circumferential bulge so that the radial compression of said hydrodynamically lubricated packing element between the specified configuration of the inner surface of the specified circular packing box seal and decree this rotatable sealing surface provides the condition of the deformation of the seal with essentially two-sided symmetry of the indicated dynamic and static bulges, and the specified one of the specified inner and outer periphery, including the specified dynamic sealing bulge, saves the configuration, so that it is effectively retained the ability of the hydrodynamic wedging grease and the ability exclusion of contamination. 6. The torsion-resistant hydrodynamically lubricated rotating stuffing box-type sealing element according to claim 1, comprising: a housing defining a circular packing box having a circumferential contact surface with a seal for radially compressing said circular static sealing protuberance to thereby provide radial compression of mostly annular body compaction without causing it to twist. 7. A hydrodynamic rotating packing element of a stuffing box type for use as a seal between a static member defining a packing box having a packing having a static sealing surface and a rotating member forming a circular dynamic sealing surface comprising: (a) a substantially annular body of elastic sealing material having inner and outer circumferential peripheries, and forming a first axial end portion for contact with contaminated material and a second axial end portion for contact with the lubricant; (b) at least one dynamic sealing bulge protruding in a first radial direction from said generally circular annular body of sealing material and configured to create dynamic sealing contact with said circular dynamic sealing surface of said rotating member, wherein said dynamic sealing the bulge forms a sharp protective shoulder at the specified first axial end part for contact with the specified contaminated material and for scraping contact with the specified circular dynamic sealing surface of the specified rotating element, while the specified dynamic sealing bulge additionally forms a non-circular shoulder for hydrodynamic proppant contact with the specified lubricant for wedging lubricant into the interface between the specified dynamic sealing convex and the specified circular dynamic sealing surface of said rotating shaft ; and (c) at least one circular static circumferential sealing bulge protruding in a second radial direction from said generally circular annular body of sealing material for sealing contact with said static sealing surface, said static sealing bulge being radially opposite to said dynamic sealing bulges, while the dynamic and static sealing bulges are essentially equal to the radial deformation under radial compression between said dynamic and static surfaces, while radially compressing said generally annular body of elastic sealing material between said dynamic and static sealing surfaces, said hydrodynamic sealing element remains substantially free of torsion as a result of said radial deformation. 8. The hydrodynamic rotating packing gland according to claim 7, wherein said dynamic and static sealing bumps have each circular configuration and are located at the end portion of the contaminants of said generally circular annular body of sealing material, wherein said dynamic sealing bump defines said non-circular a shoulder that is oriented for proppant contact with said lubricant when said dynamic sealant The convexity is radially deformed by compression with said dynamic sealing surface of said rotating shaft. 9. The hydrodynamic rotating packing element of packing type according to claim 7, wherein said static element defines a packing box containing a seal, which forms a configuration of an inner circumferential sealing surface for joint radial compression of said circular static sealing convexity of said generally circular annular body of sealing material, that there is no twisting of the specified mainly circular annular body of sealing material when radial compression.