JP2702287B2 - Dynamic pressure lubricated rotary shaft seal with torsional resistance zeometry - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION This application is filed on March 13, 1992 by Lanie Dittle and Jeffrey D. Gobeli and is referred to in U.S. patent application Ser. This is a continuation-in-part application of the issue.

FIELD OF THE INVENTION The present invention relates generally to a hydrodynamic lubricated rotary shaft seal, and more particularly to a hydrodynamic lubricated rotary with protruding hydrostatically sealed interfacial zeometry that minimizes the possibility of seal torsion as a result of radial compression. It relates to a shaft seal.

BACKGROUND OF THE INVENTION In the oilfield drilling industry, Kals is a registered trademark of Calci Engineering, Inc. of Sugarland, Texas, based on the principles of U.S. Patent No. 4,610,319.
A number of dynamic pressure lubricated ring-shaped diaphragm packing type rotary shaft seals commercially available from i Seals are used. These dynamic pressure lubricated seals have been used to perform lubricant retention and decontamination in severely worn environments such as downhole oil well drilling environments and have been successful in both low and high lubricant pressure plants. doing. Currently commercially available drilling equipment includes rotary cone rock bids, mud motors, high speed calling swivels and rotary drilling heads. All references to dynamic lubrication or dynamic seals relate to seals that implement the principles of the aforementioned U.S. Patent.

FIGS. 11 to 18 relate to a prior art dynamic pressure seal which is described to help understand the differences between the prior art seal and the seal of the present invention. The most commonly used tool for a hydrodynamic rotary shaft seal is a low lubricant pressure tool, a rotary cone drill bit, which is illustrated in FIG. The cross-sectional zeometry of the drill bit seal is determined by the shape, length and relative position of the shaft interface and by the relative position of the seal groove. The result is a dynamic pressure contact zeometry that is substantially centered with respect to the static pressure zeometry of the seal. The drill bit seal defines a hydrostatic sealing surface 1, an environmental side 2 for contacting drilling fluid, and a lubricant side 3 for contacting lubricant in the lubricant chamber of the bit. The seal also forms a steep exclusion edge 4 and a hydrodynamic sealing interface 5 which contacts the cylindrical sealing surface 6 of the shaft and forms a non-circular hydrodynamic edge 7 exposed to the lubricant. Form. The corner bend 8 of the shaft places the environmental or muddy side 2 of the seal and the steep edge 4 in a laterally offset relationship. This feature substantially centers the hydrostatic sealing interface of the drill bit with respect to the hydrostatic interface. Thus, the drill bit compresses radially and does not induce twisting of the seal. FIG. 13 is a plan view of the contact area between the cylindrical dynamic pressure sealing interface 5 of the drill bit seal and the cylindrical sealing surface 6 of the shaft, that is, the "footprint", is shown in FIG.

Drill bit seal testing Field tests using a large number of samples of dynamic pressure drill bit seals and non-dynamic pressure bit seals of normal elliptical cross-section show that the dynamic pressure features and reject features of the dynamic pressure seals and their relative rotation It has been shown to be effective in reducing shaft surface wear. In a few cases, the inner surface of the dynamic pressure seal was practically not worn after the long-term test. That is, the finished surface as it came out of the mold cavity was still seen. Under the same test conditions, the standard non-dynamic seal used as a symmetrical sample showed significant grooves on the inside surface and also had a corresponding groove on the shaft sealing surface. Analysis of the data from the field tests concluded that the bit maker said that "the use of a hydrodynamically lubricated seal significantly improves the life and reliability of hermetically sealed bearing lock bits."・ Used in drill bits.

During the development of the hydrodynamic drill bit, a wide range of hydrodynamic drill bit seal zeometry was used to control the shape and location of the hydrodynamic sealing interface, the nature of the hydrodynamic interface contact pressure profile and the fit of the entire cross-section during packing gland. A moving finite element analysis was performed. Predictions for bit seal film thickness and interfacial contact pressure are due to the fact that seal lubrication induced by the "wedge" effect is adequate in typical operating conditions and that the interfacial pressure distribution increases sharply near the rejection edge. Demonstrated very desirable.

The geometry of the interfacial contact area of the bit seal was also measured experimentally under various pressure and temperature conditions found during actual operation. The measurement results show that the bit seal interface contact area or “footprint” indicated by “F” in FIG. 13 gradually expands with an increase in temperature and compressive force, but the waveform of the dynamic pressure edge A has a very high height and shape. And the sharp exclusion edge "B" of the seal remained in contact with the shaft and showed no significant blistering or tendency to separate from the shaft. Footprint width measurements were closely correlated with the finite element analysis width estimates.

Finite element analysis of the bit seal showed that the cross-sectional zeometry of the bit seal as molded at the average wave height was relatively stable upon compression and did not have a large tendency to tilt in one or the other direction. The inherent stability of this design is related to the fact that the inner hydrostatic zeometry is substantially axially centered with respect to the outer hydrostatic zeometry.

This dynamic bit seal is relatively stable in low or zero pressure applications, but has a mud resistance of 500-1500 psi.
It may not be suitable for high pressure applications, such as motor sealed bearing lubricant retention operations. This is because the dynamic pressure sealing zeometry is not directly supported by the wall of the seal packing retainer. If the bit seal is exposed to high lubricant pressure, the seal will be pushed toward the side wall tie opposite the gland. Initial contact with the packing gland will occur at a diameter substantially larger than the sealing interface with the shaft. Static pressure acting on the unsupported area between the wall contact surface and the shaft contact surface will exert a huge force on the inner diameter of the seal. As a result, the lubricant pressure will push the inner diameter portion of the seal toward the opposite packing gland, which will be subject to abnormal distortion. Such large distortions in hydrodynamic zeometry prevent hydrodynamic lubrication of the interface, and therefore will result in rapid seal failure as a result of direct frictional contact, melting by frictional heat and rapid extrusion by elastomer softening.

Universal Dynamic Pressure Seals Universal dynamic pressure seals have also been developed within the spirit of the aforementioned US patents. In FIGS. 12 and 12A showing the prior art, FIG. 12 shows a state where the seal is compressed in the packing gland, and FIG. 12A shows an uncompressed state of the seal. In FIG. 12, a typical universal hydrodynamic rotary shaft seal E is installed in a circular housing groove, which radially compresses an elastic circular sealing element against the cylindrical sealing surface of the shaft. It is sized to bear against and thus provides a hydrostatic seal to the housing, and a dynamic pressure seal to the rotating shaft, similar to conventional throttle packing seals such as O-rings. When the shaft begins to rotate, the hydrostatic seal is stationary with respect to the housing and forms a hydrostatic seal interface with the housing, and the seal-shaft interface is the hydrodynamic seal interface.

As shown in FIGS. 12, 12A and 13, the inner surface D of the hydrodynamic seal hydrodynamically lubricates and eliminates environmental pollutants from the hydrodynamic seal / shaft interface area. It has a zeometry that results in the life of the seal. The inner circumference zeometry of the dynamic pressure seal includes an edge A that fluctuates axially in a wavy fashion on the lubricant side, and a steep exclusion edge B on the environmental side C. When the relative rotational movement of the shaft is started, the lubricant side L having a shape G that gradually concentrates in the axial direction
13 produces a dynamic pressure wedge action (action by the vertical component Vn of the rotation speed V), and this action introduces a lubricant film between the seal E and the shaft H as shown in FIG. This lubricant film physically separates the seal from the shaft, thus preventing the typical dry-friction type wear seen with prior art non-hydrodynamic squeeze-packing type seals, and thereby the seal and shaft. The service life of the corresponding shaft is extended, and higher working compression is possible. The straight steep edge B at the seal / shaft interface on the environmental side C does not produce any dynamic pressure wedge action without axial fluctuations and thus acts to exclude contaminant particles from the seal / shaft interface. In many types of rotating devices, a slight axial movement occurs due to the flexibility of the components and the various internal gaps. The steep corner B of the non-dynamic side edge of the sealing interface is generally known as the reject side or reject edge and rejects contaminants by the action of scraping off the axis during such axial movement. Thus, when relative axial motion occurs between the shaft and the seal, the contaminants that are deposited, typically abrasive, are scraped off the sealing surface of the shaft, thereby removing the contaminants from the hydrostatic sealing interface. You. In addition, the dynamic wedge effect created by the interaction of the lubricant with the dynamic pressure edge A results in a controlled lubricant pumping action from the lubricant side of the dynamic pressure interface to the environmental side, resulting in minimal lubrication. This will cause a leak rate. Such leakage or lubricant film motion tends to eject contaminant particles from the interface to the contaminant side of the seal, regardless of any sealant entering the dynamic pressure sealing interface through the exclusion edge B. .

The cross-sectional shape of the hydrodynamic rotary shaft seal used in the drill bit differs significantly from the cross-sectional shapes marketed for other applications. The space in the drill bit is severely limited in consideration of mechanical strength, and the cross-sectional geometry of the hydrodynamic bit seal is determined by the shape and length of the shaft interface, the geometry and relative position of the seal groove. Such a constraint on the seal results in a zeometry of the hydrostatic sealing interface substantially centered with respect to the hydrostatic sealing interface. In this regard, the cross section of the bit seal is substantially different from the cross section of the standard general purpose seal seal production line. Unlike the bit seals, this standard type seal is designed for high pressure and low pressure. These universal seals have a hydrodynamic sealing interface that is offset from the hydrostatic sealing interface. Unlike the bit seals that are designed for specific low pressures, universal dynamic pressure seals are designed to be used for high and low pressures. For this reason, the cross-sectional shape of the general-purpose seal is substantially different from that of the bit seal. The steep exclusion edge is preferably located at the environmental end of the seal. The location of the reject edge on the universal dynamic pressure seal is determined by the need to support from the packing gland wall in high pressure applications. This configuration is such that when high lubricant pressure conditions exist, the pressure difference between the lubricant and the contaminant causes the reject side of the seal to be pressed against the packing retainer wall as shown in FIG. Determined by the facts. If the exclusion edge is not at the end of the environmental side of the seal, the seal will not be able to perform dynamic pressure action due to significant distortion, as described above for the high pressure bit seal case.

The design of the universal dynamic pressure seal is such that the steep exclusion edge is located at the end of the environmental side of the seal, and the shape of the environmental side of the seal is substantially the same as the shape of the packing wall on the environmental side Therefore, the universal seal is well supported against the lubricant pressure at all points except the gap I existing between the housing and the shaft. This gap is often called the extrusion gap and is designed to be very small so that the synthesis of the sealing material can bridge this gap and resist deformation caused by dynamic pressure.

DESCRIPTION OF THE PROBLEM The universal dynamic pressure seal shown in FIG. 12 is required to hold down lubricant in an abrasive material environment (as evidenced by its widespread use as a high pressure mud motor seal). Works very well, but often exhibits premature wear when used to hold a lubricant under non-compression or low pressure in an abrasive environment.

The most common low pressure application of universal dynamic pressure seals is the sealed bearing assembly of oil well drilling mud motors, where such bearing assemblies are used to drill into hard soils and rock formations, This is a difficult shaft sealing application.

The low pressure seal on the mud motor must operate under the following conflicting set of conditions.

1. High fluctuation level of lateral shaft deflection, 2. High temperature environment due to geothermal, bearing and seal generated heat, 3. Extremely abrasive drilling fluid environment, 4. Limited lubricant tank volume, 5. Severe radial direction And axial space constraints, 6. Misalignment of fixed shaft and housing, 7. High level vibration, 8. Axial movement due to mechanical internal clearance and elasticity of components, 9. Axial movement of seal due to pressure balance .

Low pressure drilling mud motor seals must be capable of operating with a low leakage rate and a reliable long life under these adverse combinations of conditions. In addition, the high pressure seal of the mud motor has the above problems,
Must withstand 0-1500 psi lubricant pressure.

The mud motor is a forward moving hydraulic motor that is arranged at the lower end of the drill string and generates the rotating power of the drill bit. The motor is driven by circulating drilling fluid, which is also used to drain drilling cuttings from the well.

The entire mud motor consists of three main subassemblies. That is, it includes a hydraulic motor, a universal joint and a bearing assembly. The circulating drilling fluid drives the rotor of the motor.
A universal joint transmits this rotational movement from the rotor to the axis of rotation of the bearing subassembly, to which the drill bit is threadably connected. Weight is transmitted from the drill string to the drill bit via the thrust bearing of the bearing subassembly.

As the bit rotates, the bit impacts the formation and fractures under the effect of the added weight, so that the bit weight is concentrated by the bit cutting structure. The radial bearings of the bearing subassembly direct and guide the bit relative to the drill string.

In steering systems, a vent housing is used between the motor subassembly and the bearing subassembly. Direction control is performed by rotating the drill string to momentarily orient the vent housing in the desired direction of travel. Straight drilling is performed by continuous rotation of the drill string. The vent housing applies a passing lateral load to the radial bearings of the bearing subassembly in both straight and directional drilling modes, and in conjunction with frequent variations in axial load, Contributes to shaft deflection fluctuation.

All housings of the bearing assembly of the mud motor are filled with bearing lubricant, which is held at both ends of the bearing housing structure by rotary sealing elements. The lubricant is pressure balanced against the pressure of the drilling fluid, and many designs balance the lubricant against the pressure in the drill string holes. This pressure equalization is performed by a pressure equalizing piston, which is in a sliding sealing relation to the housing and in a rotary sealing relation to the shaft.
The rotary seal of the pressure balancing piston is one of two locations where universal dynamic pressure seals are commonly used as low pressure seals in sealed bearing assemblies for drilling mud motors.

After the drilling fluid has passed through the hollow shaft of the sealed bearing subassembly and then through the drill bit jet and into the well of the well, its pressure is reduced to a level about 500-1500 psi below the drill string bore pressure. drop down. Since the lubricant pressure in the bearing assembly is balanced to either the drill string pressure or the analass pressure, one of the rotary seals
One must include a 500-1500 psi pressure drop between the lubricant and the environment. Since the lubricant is usually balanced against the drill string, a high pressure seal is generally used towards the lower end of the assembly.

As experience has shown, it is desirable to provide a barrier fluid and a rotary barrier seal to protect the high pressure seal and the surrounding mechanical structure from the abrasive drilling fluid environment. The barrier seal structure creates a clean lubricant environment on both sides of the high pressure seal, the second of two locations where the universal dynamic pressure seal is used as a low pressure seal in the sealed bearing assembly of a drilling mud motor. I do.

If a barrier seal is not used, the high pressure seal and the surrounding mechanical structure are subject to abrasive wear, which degrades the performance of the high pressure seal in various ways. Abrasive material in the extrusion gap present between the housing and the shaft causes significant wear to the shaft and corresponding holes in the housing,
This reduces the extrusion resistance of the high pressure rotary seal, as it increases the extrusion gap. Abrasive shaft wear takes the form of local grooves. When relative axial movement occurs between the housing and the shaft as a result of the gap in the relative internal assembly and the elasticity of the support components, the tip of the high pressure seal can be broken by climbing over the groove. Also, if a barrier seal is not used, a few drilling fluid media can chemically corrode the material of the high pressure seal.

Generally, the barrier seal is mounted in a floating pressure balancing piston that has a sliding sealing relationship with the housing bore and a rotating sealing relationship with the shaft. The purpose of this floating piston is to pressure balance the barrier lubricant to the environment. Barrier lubricant is the lubricant trapped between the barrier seal and the high pressure seal.

Mud Motor Dynamic Pressure Seal Wear Study Investigations of low pressure dynamic seals used in a number of different mud motor sealed bearing subassembly structures show that premature wear of low pressure seals occurs frequently. . When described in one document, groove-type attrition appeared on the environmental side of the inner periphery of the low pressure seal after less than 70 hours of operation, but both seals were still effective. Obviously, this type of abrasion is caused by the ingress of abrasive material into the hydrodynamic sealing interface, i.e., abrasive particles entering the hydrodynamic sealing interface beyond the exclusion edge of the seal. One low pressure seal was worn after operating in the mud motor barrier seal for over 200 hours but was still effective. A number of other low pressure seals were seen that, while having good performance, were completely and smoothly abraded on the inner surface, leaving virtually no indication of the first axially fluctuating shape.

Given much evidence regarding premature wear of general purpose dynamic seals used at low pressure in mud motors, tests were conducted to study this problem under controlled laboratory conditions. This test involved a universal dynamic seal operating in the presence of a circulating agent under minimal pressure on a finely polished tungsten carbide surface. The seal applied a momentary lubricant pressure prior to rotation sufficient to orient the seal accurately relative to the environmental aspect of the gland. The environmental aspect of the seal was exposed to an abrasive grinding fluid containing sand. During this test, the lubricant temperature was kept within the temperature range expected for mud motor operation.

During the test process, the assembly was disassembled approximately every 24 hours to look for signs of wear. Prior to reassembly, the seal and shaft were thoroughly cleaned. Momentary lubricant pressure was applied to orient the assembled seal against the environmental packing gland.

Despite the daily cleaning of the sealed interface and the low twist of the shaft, the seal began to show wear bands when operated for about 100 hours. At about 140 hours, the width of the attrition band increased significantly. In about 160 hours, the torque suddenly increased,
The torque value has more than doubled. A decrease in leakage rate was seen with increasing torque. At about 170 hours, the torque had almost dropped to its initial value, had been running at this level for about 24 hours, and then began to increase. At about 200 hours, the test was terminated due to the high torque level. At the end of the test, the seal was still effective, but exhibited a wear pattern as shown in FIG.
It should be noted that a circular wear groove J is formed at the hydrodynamic sealing interface on the contaminant side C of the seal.

From about 110 hours to the end of the test, a progressive increase in the attrition pattern was seen at this point of the sealed interface. About 100
The wear pattern of the shaft was finally visible over time, but could not be felt. From about 200 hours, the shaft wear pattern showed a matte texture that could be easily felt.

It was initially thought that the matte structure of the shaft wear pattern or the abraded material that had entered was responsible for the increased torque.
This idea was violated, however, by the fact that the worn seals, possibly after cleaning, had a high torque when operated on a new shaft. Similarly, the new seal did not exhibit high torque when operated on the shaft wear pattern. It was concluded that the increase in torque that occurred about 150 hours into the test was due to loss of hydrodynamic lubrication due to seal wear. This conclusion is supported by the increased torque and reduced leakage rate of the stern.

The seal was effective at the end of the 200 hour test, but the barrier seal did not operate for this long in actual use due to more demanding conditions such as more dynamic operation and disassembly and cleaning every 24 hours Was.

In the above tests, severe degradation of dynamic pressure lubrication and corresponding increase in seal wear rate, as well as sharp increase in torque, are significant. Many of the low pressure seals in mud motors are mounted in floating pistons. Outer sliding seal friction occurs at the sealing interface between the only existing piston and housing. The outer sliding seal is an O-ring or an assembled seal as manufactured and sold under the trademark PolyPack by Parker Seal. The sudden increase in torque causes the piston and sliding seal to rotate within the housing. Such rotation is very destructive to the outer seal because the outer seal is in dry frictional contact with the housing bore. This resulted in severe wear of the outer seal, which was reported by several users and witnessed by the applicant. The sudden increase in torque results in a dynamic pressure seal spin in the gland and wear of the seal and gland due to dry sliding contact. To prevent this, many glands are subjected to grid blasting to increase frictional resistance for static retention of the seal.

If there is an eccentricity between the shaft and the gland, the radial seal compression decreases at about one half of the seal circumference and increases in the other half. In order to maintain a fluid-tight seal in such an eccentric state, sufficient pre-compression must be performed to ensure a sufficient level of compression in the offset state. At the same time, the maximum degree of compression seen in the offset state must not adversely affect the sealing performance. For this reason, it is desirable that the seal be able to withstand a wide range of compression.

Finite element analysis of a universal dynamic pressure seal Although finite element analysis was not initially used in the initial design of the universal dynamic pressure seal of the present invention, at a later date in the laboratory, a portion of the seal was Finite element analysis was used extensively to investigate the case where the contact pattern on the shaft showed some friction with the rotating shaft. Since that time, other examples of shaft lubricant contacts have been found in low pressure seal samples taken from drilling mud motors used in oil wells.

Finite element analysis was performed on a typical general-purpose static pressure seal, and the deformation and interface contact pressure of the seal obtained under various temperature conditions, compression conditions, and lubricant pressure conditions were evaluated. This analysis showed that the universal seal became increasingly distorted under increasing radial compression levels, and in some cases the lubricant end of the seal contacted the shaft. FIG. 15 for the prior art shows a finite element analysis transfer plot of an incompressible universal seal modeled at half-wave height at 11% compression and a 180 ° F. temperature rise. The wavy line M indicates the undeformed state of the seal. Solid line N shows the expected deformation of the seal in radial compression. This displacement plot shows that in the modeled condition, the seal twists until the inner circular corner F at the lubricant end of the seal contacts the axial surface Q. A low pressure seal / shaft contact pattern exhibiting such contact was seen in laboratory test samples as well as on mud motor low pressure seals returned from the field.

Also, finite element analysis of uncompressed seals shows other undesirable trends. That is, due to the torsion of the seal during packing, the steep extrusion edge B of the seal lifts off the shaft even at relatively low initial pressure. Such a lifting tendency decreases as the degree of compression increases. No.
FIG. 18 is an enlarged cross-sectional view of the exclusion rim of a moving plot of an uncompressed universal seal modeled at half-wave height at 11% compression and a 180 ° F. temperature rise as in FIG.
FIG. 18 shows a state in which the exclusion edge R has been lifted from the rotation axis Q as a result of the seal torsion induced by compression. As a result, the seal has an S-type that gradually concentrates on the environmental side of the dynamic pressure seal interface T. The loosely concentrated shape of the seal is reflected in the interface pressure profile predicted by finite element analysis.

Ideally, the contact pressure profile on the environmental side of the hydrodynamic sealing interface should rise sharply at the steep edge of the seal, because the steep edge is for producing a scraping action. If a universal hydrostatic seal is used for non-pressure lubricant, there is no scraping action. Instead, local seal zeometry takes the form of a gradual concentration with the aforementioned shaft, which is ideal for trapping abrasive material during axial movement of the shaft. Due to the flexibility of the bearings and other mechanical clearances and components, there is micro axial movement of the shaft in the mud motor sealed bearing assembly. As a result of raising the steep edge of the seal and locally concentrating zeometry of the seal, the axial movement drives the abrasive material into the hydrodynamic sealing interface.
This results in seal and shaft wear.

The degree of twisting of the seal during packing gland is related to the local height of the dynamic pressure wave. The steep edge slope and corresponding lift are large at the low point of the wave and small at the high point of the wave. This means that the last contact point on the environmental side of the dynamic pressure sealing interface fluctuates in the axial direction as a function of the dynamic pressure introduction wave height. Unfortunately, this phenomenon limits the dynamic pressure action of the exclusion edge, which causes the polluting particles to wedge into the sealing interface and premature wear of the seal in the absence of axial axial movement. And thus limit the operating life of the seal. This conclusion is supported by wear seals obtained from laboratory tests where there was no axial movement.

Also, a finite element analysis of 11% compression and 180 ° F. as described above was performed at a differential pressure of 100 psi as shown in FIG. From this load, it was expected that the pressure on the lubricant side would cause the seal to stand upright and abut against the lubricant wall on the environment side. This brought the sharp edge into contact with the shaft as expected, resulting in a sharp contact pressure profile and the expected scraping action on the environmental side of the sealing interface as expected. Also, the seal upright effect of the lubricant showed a desirable effect of preventing contact with the shaft at the lubricant end of the seal.

Also, as shown in FIG. 17, a general purpose seal was modeled at a compression ratio of 15% and a temperature difference of 180 ° F. At this compression level, virtually the entire inner part of the seal was made flat with respect to the axis. Also this seal model is 100psi
When the differential pressure was applied to the lubricant side, it remained practically flat with respect to the shaft. Such extreme distortion of hydrodynamic zeometry severely inhibits or completely eliminates the desired hydrodynamic pumping action of the seal. This is because the dynamic pressure edge G is actually separated from the lubricant by the contact of the circular edge F with the shaft.

From this finite element analysis, up to 12% compression can be used in the presence of 100 psi or higher lubricant pressure, but the initial compression of a universal seal is limited to 9% or less in the absence of lubricant pressure. It was concluded that it had to be.
Based on the results of this finite element analysis, Calci Engineering, Inc. has begun to recommend a degree of compression of 7-1 / 2%. Clearly, a seal that can withstand higher compression levels is desirable. Such seals are better able to withstand misalignment, runout, manufacturing tolerances, compression hardening, and wear.

In principle, smaller cross-section seals must undergo higher initial compression than larger cross-section seals. This is because a certain degree of misalignment, dynamic runout, tolerance, wear, etc., corresponds to a higher percentage as the seal cross section size is smaller. Due to the compression limits of current seal designs, there are corresponding limitations on the minimum effective seal cross-sectional size that can be manufactured.

Summarizing the results of the finite element analysis, the torsional action of the pressureless seal, which occurs when the seal is subjected to radial compression, impairs the desired scraping action of the exclusion edge and reduces the seal / shaft interface on the contaminant side of the seal. An undesirable dynamic pressure action is introduced therein to pump the abrasive material into this sealing interface. The wear of the seal due to the presence of such abrasives and the diminished lubrication drainage limits the operating life of the seal.

There is no lifting tendency of the exclusion edge in the reduction seal. This is because the lubricant pressure presses the seal against the side wall opposite the packing gland, which reorients the seal and presses the reject edge toward the axis. The rolling seals are not subject to the torsional action during the pressureless seal, but both seals are severely distorted at high compression levels, with a ban on dynamic pressure action. Higher compression levels than currently available would be desirable.

Field tests of normal hydrodynamic seals operating under low fluid pressure conditions and high mechanical compression have shown excessive wear after a relatively short period of time, and these seals are shown in FIG. It is believed that the displaced edge of the seal is now distorted under mechanical compression enough to lift up from the sealing surface of the shaft and allow abrasive material to enter therebelow.

If the current hydrodynamic shaft seal is used in applications where the lubricant pressure is different from the pressure in the contaminant environment, this differential pressure will hold the seal against one wall of the gland (as described above). In addition, the exclusion edge of the dynamic pressure lip is maintained in a perfect circular shape to prevent the position of the exclusion edge from fluctuating in the axial direction. This is not the case if the lubricant and the environment are at the same pressure. The axial width of the seal groove is naturally larger than the compressed width of the elastomer seal to accommodate the thermal expansion difference between the seal elastomer and the seal groove, and the seal moves forward and backward within the limit of the gasket. Fluctuate. Such an effect is referred to as “meandering” during packing holding. This meandering has the effect of bending the seal's displacement zeometry against the rotational speed of the shaft, such that the rotational speed of the shaft causes environmental pollutants to impinge on the exclusion zeometry and cause its wear.
Thus, if meandering during seal packing is prevented, the ultimate life of the seal will be improved in pressureless applications.

Also, the ultimate life of current dynamic pressure shaft seals is limited by the compression set (permanent deformation) of the elastomer at high temperatures. It has been observed that much of this compression hardening occurs in the dynamic pressure lip of the inner portion of the seal, rather than the entire rectangular body of the seal. There are two reasons for this phenomenon. First, the lip is at the point of self-generated heat. That is, at the shearing point of the lubricant at the sealed interface. Second, because the lip is much smaller than the body of the seal, the lip compresses more than the seal body due to a synthetic imbalance. Thus, if the compression is distributed evenly along the body of the seal rather than being concentrated in the protruding lip, the compression set resistance of the seal will be improved.

SUMMARY OF THE INVENTION The present invention is directed to a hydrodynamic throttle packing type rotary shaft which solves the above-described problems with respect to a general-purpose dynamic pressure rotary shaft seal which operates in a severe wear environment even in a state of receiving or not receiving a lubricant pressure. It is about a seal.
The seal of the present invention has an axially varying cross-sectional shape on the hydraulic side of the dynamic pressure seal interface for dynamic pressure lubrication of the dynamic pressure seal interface, and the environment of the dynamic pressure seal interface for the purpose of eliminating pollutants. It has a steep, axially stable corner exclusion that is straight to the side. More specifically, the dynamic pressure seal of the present invention relates to a commercially available dynamic rotary shaft seal manufactured and sold by Calci Engineering, Inc. of Sugarland, Texas under U.S. Patent No. 4,610,319. According to the present invention, the dynamic pressure seal of the present invention is provided to provide mechanical support for lubricant pressure acting along the dynamic pressure area of the seal to establish a cross-sectional zeometry compatible with the high pressure lubricant retaining seal. Has its exclusion edge located at one end of the cross-sectional zeometry.

The present invention also dramatically improves the performance of the universal hydrostatic seals described above, especially when used in non-pressure or low pressure lubricant retaining structures. Such an improvement is implemented by eliminating the tendency of the reject lip to lift as is presently seen. Such an improvement in the performance of the reject lip results in a substantial increase in the life of the seal when used in an abrasive environment. Improved rejection performance reduces the chances of severe torque fluctuations that would induce spinning of the seal in the gland. If this improved seal is mounted in a floating piston structure, it reduces the tendency of the piston to spin in the bore.

The present invention also extends the life of the seal by increasing the initial compression level as much as possible without compromising the seal dynamic pressure lubrication characteristics. Such additional compression makes the seal far less susceptible to compression hardening, wear, mechanical misalignment, dynamic runout and manufacturing tolerances.

The present invention also reduces the interfacial contact pressure near dynamic pressure lubrication zeometry, which makes dynamic pressure lubrication more fully effective at low speeds.

The present invention also improves the compression hardening tendency of the seal to some extent by redistributing the compression along a larger portion of the seal.

The present invention also achieves all of the above advantages by providing a simple, compact, protruding sealing zeometry at the non-dynamic pressure sealing interface. The reaction to the compressive load occurs without any difference, which results in a corresponding torsion of the zeometric sealing section. The improved seal of the present invention can be manufactured from existing molds with minor tool changes. Also, the seal of the present invention requires a small amount of material per part to be manufactured and is fitted in the same gland as the existing universal hydrodynamic seal, thus requiring retrofitting to change existing equipment. Absent.

SUMMARY OF THE INVENTION The present invention provides a hydrodynamic rotary seal having a hydrostatically projecting interfacial zeometry that substantially duplicates and mirrors the hydrodynamically projecting interfacial zeometry and a cross-sectional zeometry of adjacent portions. In such a configuration, the overall cross-sectional shape of the dynamic pressure seal becomes closer to symmetrical, which significantly reduces the tendency to twist during compression. This is because the cross-sectional distortion caused by the compression with respect to the static pressure interface becomes almost the same as the distortion due to the compression at the dynamic pressure interface. This symmetrical seal configuration is not prone to twisting of the seal when the seal is subjected to radial compressive forces in the gland.

In a laboratory, a standard general-purpose dynamic pressure seal and a dynamic pressure seal having a static pressure protruding interface zeometry according to the present invention are provided.
The test was carried out in an abrasive drilling mud environment while maintaining a lubricant pressure of 15 psi. These seals were tested on separate parts of the same coaxial surface under the same run-out, the same gland and the same temperature and speed conditions. The test of the standard universal seal is described in detail in the "Problem Disclosure" section of this specification. Briefly, the universal seal was disassembled for inspection approximately every 24 hours, then washed and reassembled. This seal shows abrasive wear in the form of a groove at 110 hours, a dramatic increase in torque in about 160 hours,
As a result, a decrease in lubricant leakage was observed. In contrast, the seal with the hydrostatic lip of the present invention did not require daily interfacial cleaning and was disassembled in 144 hours without any abrasion. After 508 hours, the test was terminated and the seal and shaft were carefully inspected. The seal also had no signs of shaft wear.
The torque was low and steady during the test. This test confirmed that the dynamic pressure seal of the present invention performed much better than the standard general purpose dynamic pressure seal.

BRIEF DESCRIPTION OF THE DRAWINGS Hereinafter, the present invention will be described with reference to embodiments shown in the drawings, but the present invention is not limited to these embodiments.

In the accompanying drawings, FIG. 1 is a partial sectional view of a housing including a dynamic pressure rotary shaft seal and a rotary shaft structure according to the present invention.

2 and 2A are partial cross-sectional views of a dynamic pressure seal according to a preferred embodiment of the seal of the present invention. FIG. 2 shows the compressed state of the seal during packing press, and FIG. 2A shows this seal. 3 shows the uncompressed state of

FIG. 3 is a partial sectional view of a dynamic pressure seal according to another embodiment of the present invention.

FIG. 4 is a partial sectional view showing a structure in which a dynamic pressure seal according to the present invention is disposed in a seal groove in a housing, and FIG. 4A shows an uncompressed state of the seal.

FIG. 5 is a partial sectional view of a dynamic pressure seal according to still another embodiment of the present invention.

FIG. 6 is a partial cross-sectional view showing another embodiment of the hydrodynamic seal of the present invention compressed in a seal groove having an outer tapered support surface, and FIG. 6A shows the seal in an uncompressed state. FIG.

FIG. 7 is a partial sectional view showing a compressed state in which the seal of the present invention is arranged in another housing, and FIG. 7A shows an uncompressed state of the seal.

FIG. 8 shows the seal of the present invention compressed in yet another housing, and FIG. 8A shows the seal in an uncompressed state.

FIG. 9 shows the seal of the present invention compressed in yet another housing, and FIG. 9A shows the seal in an uncompressed state.

FIG. 10 is a view showing a radially compressed state of another embodiment of the seal of the present invention, and FIG. 10A is a view showing its uncompressed state.

FIG. 11 is a partial sectional view showing a typical prior art dynamic pressure drill bit seal.

FIG. 12 is a partial cross-sectional view showing a state in which a general-purpose dynamic pressure seal according to the prior art is placed in a compressed state in a sealing gasket and is engaged with the rotary shaft by dynamic pressure sealing, and FIG.
The figure shows a non-compressed state of this seal.

FIG. 13 is a developed view showing a dynamic pressure wedge action and a pumping action of the dynamic pressure seal.

FIG. 14 is a partial cross-sectional view of a prior art dynamic pressure seal showing an abrasion pattern due to entry of an abrasive from the contaminant side.

FIGS. 15, 16 and 17 are finite element analysis diagrams of the prior art dynamic pressure seal, in which the uncompressed state is indicated by a broken line and the radially compressed state is indicated by a solid line.

FIG. 18 is a partial cross-sectional view showing the exclusion edge of the seal tilted or twisted and lifted from the rotating shaft in response to radial compression under a low pressure differential.

FIG. 19 is a sectional view of a dynamic pressure seal according to another embodiment of the present invention.

FIG. 20 is a partial sectional view of a dynamic pressure seal according to another embodiment of the present invention, showing a non-compressed state.

FIG. 20A is a partial cross-sectional view showing a state where the seal of FIG. 20 is disposed in a seal groove and a radial compressive force is applied between a rotating shaft and a groove wall.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Referring to FIGS. 1, 2 and 2A, an entire hydrodynamic lubricating rotary shaft assembly is shown at 10 which includes a housing 12 from which the rotary A shaft 14 extends. The housing defines an inner seal groove, sheet or packing gland 16 in which a hydrodynamic rotary seal 18 is disposed, which seal 18 is constructed in accordance with the principles of the present invention and is shown in FIGS. 2 and 2A. This is illustrated in detail in the figure. 2A shows the radially uncompressed cross-sectional shape of the improved dynamic pressure rotary shaft seal of the present invention, whereas FIG. 2 shows the rotary structure disposed in a seal groove and rotating. 2 shows the cross-sectional shape of a radially compressed seal between a shaft and a radially outer wall of the seal groove. The radial cross-sections in FIGS. 2 and 2A are representative of the average interfacial contact width of the hydrostatically sealed interface and are disclosed in US Pat. No. 4,610,319.
This is taken at a circumferential position corresponding to the midpoint of the wave height of the zeometry which produces the lubricating wedge effect described in the above item.

From a global orientation point of view, the seal end oriented toward the lubricant is face 20. This surface 20 can differ from the illustrated flat surface within the spirit of the invention. The end of the seal, which is oriented towards a severely worn or contaminated environment 22, is a face 24. This surface can also differ from the illustrated flat surface within the spirit of the invention.

According to the invention, a circular dynamic pressure sealing element 18 is arranged in a circular sheet or sheet groove 16 and is compressed against a relative rotational surface 28 of the rotating shaft 14. When the seal is installed in a circular seal groove or seal sheet, a circular radially protruding hydrostatic sealing lip or protrusion (hydrostatic circumferential seal compressed portion) 30 is compressed against the opposing surface 32. . The inner peripheral surface of the circular sealing element 18 is provided with an inner peripheral projection (dynamic pressure peripheral sealing projection) 34, which defines a dynamic pressure sealing surface 36, which is a generally circular shape of the shaft 14. Is compressed against the relative rotation opposing surface 28 which forms a sealing surface of. Static between the static pressure sealing tip (hydrostatic circumferential seal compressed part) 30 and the opposing surface 32 and between the dynamic pressure sealing lip (dynamic pressure peripheral seal projection) 34 and the relative rotation opposing surface 28 of the rotating shaft Hydrodynamic sealing element with a radial pressure sufficient to maintain a liquid-tight seal at the sealing surface
18 is retained. A dynamic pressure seal is used to separate the lubricant in the seal lubricity chamber 38 from the contaminant fluid of the environment 22 and prevent mixing of the lubricant with the contaminants in the environment.
In the known drilling industry, the contaminated fluid is a drilling fluid called "drilling mud", which contains highly abrasive particulate matter in a fluid vehicle.

The ring dynamic pressure sealing element is constructed of any suitable sealing material, including elastomeric or rubber-like sealing materials and various polymeric sealing materials.

The inner radial projection (dynamic pressure peripheral seal projection) 34 and the outer radial projection (static pressure peripheral seal compressed part) 30 form tapered surfaces 40 and 42, respectively, which are respectively opposed to the lubrication side of the seal. The tapered surface 40 is non-circular, while the tapered surface 42 is circular. These tapered surfaces 40 and 42
Has a radius of curvature that gradually concentrates on the shaft 14 and the groove surface, respectively. The tapered surface 40 and the inner cylindrical dynamic pressure sealing surface 36 are joined by a radius of curvature portion 44. The tapered surface 40 intersects at an intersection 48 at an acute angle to the cylindrical surface 46. The tapered surface 40 and radius of curvature portion 44 form a shape that is gradually concentrated on the lubricant side of the hydrodynamic sealing surface, which causes hydrodynamic migration of the lubricant film at the hydrodynamic sealing surface as described in U.S. Pat.No. 4,610,319. . The zeometry of the tapered surface 40 can be any of a number of suitable shapes that, within the spirit of the present invention, result in a gradually concentrated shape on the lubricant side of the hydrodynamic sealing lip or projection. This gradually converging shape allows the lubricant film to be pressed into the dynamic pressure sealing interface by a wedge effect in a dynamic manner, so that when a relative rotational motion occurs between the lubricant and the seal, the seal inner peripheral surface is formed at an arbitrary position. It is arranged at the position.

The circular corners or exclusion rims 50 are used to remove contaminants from the rotating shaft in the event of axial relative movement of the shaft and seal without corresponding environmental and dynamic pressure effects corresponding to the relative rotational movement of the shaft and seal. Figure 4 shows a sharp edge zeometry that acts to scrape off. The seal thus eliminates abrasive substances present on the contaminated side of the seal from entering the hydrodynamic interface.

The structure according to the invention shown in FIGS. 2 and 2A causes the sealing material to extrude and other breaks when subjected to the hydrostatic action of the lubricant pressure acting between the hydrostatic and hydrostatic sealing faces. The present invention relates to a general-purpose general-purpose dynamic pressure rotary shaft seal formed by arranging and forming the exclusion edge 50 of the seal 18 and the environmental side surface 24 so as not to be supported by the packing holding wall 52.

The lip protrudes radially inward from the body of the seal 18 at the environmental side end 24 by combining the abrupt exclusion corner 50 with the tapered surface 40 of the zeometry, which is inclined and gradually concentrated, and the radius of curvature 44. (Dynamic pressure circumferential seal projection)
34 can be formed. The protruding lip (dynamic pressure peripheral seal projection) 34 is compressed to form a dynamic pressure sealing surface together with the relative rotation facing surface 28 of the shaft 14.

A zeometric feature of the preferred embodiment of the present invention is a hydrostatic sealing lip 30 that protrudes radially outward from the body of the seal 18 at the environmental end 24 of the seal cross section. This sealing lip (statically compressed peripheral seal compressed portion) 30 forms a static sealing surface which, when compressed, cooperates with the facing surface 32 of the circular seal groove.

In the preferred embodiment shown in FIGS. 2 and 2A,
The hydrostatic sealing lip (compressed portion of the hydrostatic circumferential seal) 30 has a cross-sectional shape substantially similar to the cross-sectional shape of the dynamic pressure lip (hydrodynamic circumferential seal projection) 34, The protruding height 54 of the compressed portion 30 is slightly smaller than the protruding height 56 of the dynamic pressure lip (dynamic pressure peripheral seal projection) 34. (The width of the hydrostatic sealing lip 30 can be varied to correspond to variations in the width of the dynamic pressure lip under given environmental conditions.) Ideally, the protrusion height 54 of the hydrostatic lip is The projection is 56 or less than the nominal radial compression, and the projection height 56 of the dynamic pressure lip (dynamic pressure peripheral seal projection) 34 is greater than 1/2 of the nominal radial compression of the seal.

Since the protruding height 54 of the hydrostatic lip 30 is less than or equal to 1/2 of the radial compressibility of the seal as shown in FIG. 2, all or almost all of the sealing surface from the circular corners 58 to 60 is provided. Faces the seal groove when the seal is compressed.
Contacted directly with 32. The close proximity and / or contact of the seal and seal groove near the cylindrical corner 58 provides stability against clockwise twisting of the seal inside the seal. In this case, the twisting action is the action seen in FIG. Static or dynamic stresses on the environmental end 24 of the seal, such as during temporary pressure fluctuations or when the contaminant edge 50 of the seal needs to quickly scrape contaminants off the axis that moves axially Such seal stability is very important when the seal is added.

The protrusion height of the dynamic pressure lip (dynamic pressure seal projection) 34 is preferably greater than half the nominal degree of compression of the seal, and the sealing surface near the inner circular corner 62 is the relative rotational facing surface 28 of the shaft 14. , The inclined hydrodynamic inlets from 48 to 44 do not press too flat against the opposing surface of the shaft 14, and therefore the large deformation of the hydrodynamic zeometry causes The desired dynamic pressure wedge action intrusion of the lubricant into the lubricant is not impeded. This means that a hydrostatic seal with a protruding hydrostatic lip according to the invention experiences a higher level of compression than a conventional hydrostatic seal. Due to the high compression level of the lip according to the invention, this seal can withstand greater mechanical misalignment, dynamic run-out, compression hardening, abrasion, allows for greater mechanical tolerances, and has a smaller cross-section seal Is realized.

In FIGS. 2 and 2A, the overall cross-sectional shape of the lip (hydrostatic circumferential seal compressed portion) 30 is substantially equal to the cross-sectional shape of the lip (dynamic pressure circumferential seal projection) 34, and therefore, when the seal is compressed, The interface contact profile of both lips and the magnitude and location of the degree of deformation are approximately equal, so that there is no tendency to overtwist the seal in the groove and an equilibrium is obtained.

The present invention virtually eliminates the tendency of the seal to twist within the groove, thereby eliminating the tendency for the sharp edge 50 to be lifted from the opposing surface of the shaft 14. This ensures that the sharp edge 50 is kept compressed against the shaft 14,
Since the scraping action can be performed, it means that contaminants from the external environment do not enter the dynamic pressure sealing interface. Such improved rejection means that the ingress of friable material into the sealing interface is greatly reduced, so that wear of the seal and the corresponding shaft surface is minimized. In this way, the life of the seal is greatly extended, as shown in the tests, and the rotational torque and the dynamic pressure leakage rate remain very uniform during the life of the seal, since there is no wear. Relatively low shaft wear means that the shaft does not need to be repainted and polished as often as in the normal case, thus resulting in substantial cost savings. According to the present invention, since the seal torque is stable, the spinning of the seal in the groove due to friction between the seal and the shaft is minimized, and as a result, the groove processing such as grit spraying currently performed is performed. It is no longer necessary.

The dynamic pressure interfacial contact pressure is reduced near the gradually concentrated dynamic pressure-introduced zeometry due to the significant elimination of the torsional tendency of the seal inside the groove. This is because at this point the seal material no longer counteracts the large torsional action of the entire cross section of the seal. Because of the reduced contact pressure at this point, dynamic pressure lubrication is complete and efficient even at low speeds.

By using two opposing lips, one hydrostatic lip and one hydrodynamic lip, the actual degree of compression of the hydrodynamic lip is lower than in the case of conventional dynamic pressure seals of the type currently used. As a result, the compression pressure resistance is improved. Because most of the compression pressure occurs at the dynamic pressure lip, which is close to the heat generated at the dynamic pressure sealing interface.

Sealed mud motors are now increasingly being used in "hot" holes, which often exceed the operating temperature limits of elastomers used in hydrodynamic seals. Elastomers currently used in dynamic pressure seals are selected primarily based on high frictional resistance and typically have a maximum operating temperature range of 250 ° F (120 ° C) to 320 ° F (160 ° C). An elastomer such as a fluoroelastomer having a higher temperature limit but lower friction resistance can be used for the dynamic pressure seal of the present invention.
This is due to the excellent torsional resistance sealing where the abrasive material is excluded from the sealing interface.

The present invention is applied to an application in which either the housing or the shaft is a rotating member and the rotating shaft is sealed to the housing. If the seal is compressed radially, the protruding lips can be located on the outer and inner surfaces of the seal cross section, and the dynamic lip is located on either the inner or outer surface of the seal. Alternatively, the protruding lips can be located at both ends of the seal cross-section if the seal is axially compressed against a relatively flat rotating surface.

In the figures, the dynamic pressure lip and the hydrostatic lip extend completely to the contaminated end of the seal, but this is not essential within the spirit of the invention. It is sufficient that these lips are arranged so as not to cause twisting of the seal in the groove upon radial compression of the seal under low pressure conditions.

In another embodiment of the invention shown in FIG. 3, the dynamic pressure seal 66 is substantially equivalent to that shown in FIG. 2A, but with a corner 64 having the radius of curvature of the seal of FIG. 2A. Steep corners 68 are used. This steep corner 68
Is to prevent the lubricant from being pumped into the hydrostatic interface when the seal is swung back and forth in the groove under fluctuating and / or reverse pressure conditions. , Thereby further reducing the tendency of the seal inside the groove to twist.

It will be appreciated that the resilient squeeze-type sealing element is an imperfect sealing device in that situation and cannot perform the sealing action without the cooperation of its packing gland.
Thus, the entire dynamic pressure sealing device consists of a combination of a packing gland and an elastic sealing element. FIGS. 4 and 4A show a compressed state and a non-compressed state of another embodiment of the seal of the present invention, respectively. As shown in FIG. 4, the gasket retaining groove 70 cooperates with the hydrostatic sealing lip 75 of the hydrodynamic sealing element 76 to set the seal in the groove 70 in the opposite direction. It prevents axial movement and also prevents meandering movement of the seal inside the groove in the uncompressed state. In this configuration, the radial projection height of the hydrostatic lip
72 is substantially increased from that shown in FIGS. 2, 2A and 3, so that this protrusion height is greater than half the initial compression of the seal and the opposite dynamic pressure It is equal to the degree of protrusion of the lip (dynamic pressure seal protrusion). Receives the hydrostatic lip (compressed part of the hydrostatic circumferential seal) 75 and reduces it to about 1/2 of the radial compression.
A corresponding groove 74 is formed in the packing gland 70 for radial compression up to that point. Compression of the static pressure lip (compressed portion of the static pressure circumferential seal) 75 thus obtained,
Deformation and deformed shape are dynamic pressure lip (dynamic pressure peripheral seal projection)
Approximately identical, the size and position of the interface contact force profile of these two lips are similar and symmetrically opposed, so that the seal is
Equilibrium is obtained without the tendency to twist significantly within 70. The similarity of the contact force profile is due to the similarity of the protruding height zeometry of the hydrostatic lip (compressed part of the hydrostatic peripheral seal) and the dynamic pressure lip (protrusion of the hydrodynamic peripheral seal) and the undeformed zeometry. Is enhanced by something similar to

When compressing the seal, its rear end 77 is
A sealing contact is made at 78 to prevent clockwise twisting of the seal due to external loads. In this case, the circular corner 78 cooperates with the cylindrical support surface 79 to provide an anti-twist reaction. If the seal of FIG. 4 is inadvertently installed in the opposite direction, the seal will protrude significantly from the groove 70, making it extremely difficult and physically impossible to set the axis, thus resulting in this carelessness. Draw attention to the appropriate settings. In addition, the zeometry of the static pressure lip (statically compressed peripheral seal compressed area) 75 and the corresponding groove zeometry provide clues to proper orientation prior to installation. When properly installed, the static pressure lip (statically compressed peripheral seal compressed portion) 75 is supplemented by the zeometry of the corresponding seal groove 74. The corresponding zeometry of seal 76 and groove 70 also constrains the axial movement of the seal against reversal of pressure. The mutually compatible zeometry of the seal 76 and the seal groove 70 also prevents the seal from meandering in the groove without applying separate lateral pressure on the seal, thus eliminating steep edge exclusion. As it is constrained, this exclusion edge does not distort against the rotational speed of the shaft,
Collisions that cause this edge wear are prevented. In addition, this mutual seal / seal groove zeometry increases the structural integrity of the seal and thus the radial compressibility of the seal as compared to other hydrodynamic seals.

FIG. 5 shows a hydrodynamic sealing element 80 according to another embodiment of the present invention.
3 is a radial cross-sectional view of FIG. The seal structure shown in FIG. 5 is a slight modification of the seal structure of FIGS. 2, 2A and 3. Static pressure lip (Compressed part of static pressure circumferential seal) 82
Is a tapered portion of the seal forming a bevel. The protruding height of this static pressure lip (hydrostatic circumferential seal compressed portion) 82 is approximately equal to half the total radial compressibility of the seal. Therefore, the degree of compression of the static pressure lip (hydrostatic peripheral seal compressed portion) 82 is substantially equal to the degree of compression of the opposite dynamic pressure lip (dynamic pressure peripheral seal projection), and the size of the interface contact profile between these two lips The locations are basically similar to each other and are symmetrically opposed so that the seal has less tendency to twist within the groove. Second
Compared with the cross-sections of FIG. 2, FIG. 2A and FIG. 3, the dissimilarities in the zeometry of the hydrostatic lip (hydrostatic peripheral seal compressed portion) and the dynamic pressure lip (dynamic pressure peripheral seal projection) are large. Although not as effective as a seal, the shape of FIG. 5 still provides the same general advantages as the previous embodiment.

Referring to FIGS. 6-10, controlled radial compression of the hydrodynamic seal is effected by the cooperation of the inner surface of the seal groove and the outer surface of the hydrodynamic seal. In each of these embodiments, the radial compression does not induce a torsion of the seal within the groove, as the radial compression results in a substantially symmetric deformation at both the hydrostatic and the dynamic pressure interfaces.

With particular reference to FIGS. 6 and 6A, which show the compressed and uncompressed states of the seal, respectively,
Defines an internal groove 92 having a tapered or frusto-conical radiating surface 94. Also, the seal 96 has a radially outward tapered surface having a steeper taper than the aforementioned tapered surface 94.
Define 98. In an uncompressed state, the seal 96 fits within the groove 92 such that its radially outer corner 100 closely engages the radially outer corner 102 of the groove 92. Thus, when the rotating shaft 104 is inserted through the seal, the seal is deformed at its dynamic pressure projection (dynamic pressure peripheral seal projection) 106 and its radially opposite portion and is substantially symmetrical to the seal. Radial compression occurs, resulting in an interfacial adhesion between the surface of the sealing lip and the opposing surface that are approximately equivalent and symmetrically oriented with respect to each other. When the seal is thus compressed, the seal no longer has a tendency to tilt or twist in the groove, so the reject rim 108 is
Optimal contaminant elimination engagement with the cylindrical surface of 04. 4 and 4A show enhanced torsional resistance and contaminant rejection properties for similar reasons. Further, in the compressed state shown in FIG. 6, the outer circular corner 99 of the seal engages the support surface 94 of the seal groove to provide a support structure for the inclination of the seal within the groove. Due to the taper of the surface 94, the radial compressive force on the seal pushes the seal toward the inner circular corner 102 of the seal groove, thus supporting the contaminant side of the seal with no additional pressure on the seal. Due to the strong seating against 103, the life of the seal is extended without the sharp exclusion edge 108 distorting against the rotational speed of the shaft and without causing this edge to wear out.

Also, the cooperative structure of the seal 96 and the seal groove 92 prevents the seal inside the groove from being installed in the reverse direction. If the dynamic pressure side of the seal is installed in the opposite direction, facing the contaminant side, the seal will be over-compressed radially, making installation of the shaft difficult or impossible. The dynamic pressure seals of FIGS. 4 and 4A are also difficult or impossible to reversely install for similar reasons.

Referring to FIGS. 7 and 7A, the housing structure 13
The zero inner groove 132 defines an inner compression surface 134, which is provided with a groove relief 136 facing the outer surface portion 138 of the seal 140. To aid in understanding this embodiment, seal 140 is shown in its radially compressed state in FIG. 7 and in its uncompressed state in FIG. 7A.
The seal shape shown in FIG. 7A is a general-purpose dynamic pressure seal which is deformed into various shapes, and the seal compression surface 134 is disposed at a position radially opposed to the dynamic pressure projection (dynamic pressure seal projection) 142 of the seal. , And has a width substantially equal to the width of the surface 144 of the dynamic pressure projection 142. As such, when the rotating shaft 146 is installed to compress the seal 140, the radially outer static pressure portion and the radially inner dynamic pressure portion 142 of the seal undergo substantially symmetrical compression and deformation. Due to the symmetric radial compression and the resulting symmetry in the interfacial contact force profile on both sides, the hydrodynamic sealing element 142 does not undergo a tilting or twisting action when it is radially compressed. Thus, this seal can undergo much greater radial compression than commercially available dynamic pressure seals. Due to the dynamic compression capability provided by the combination of these hydrodynamic sealing elements and seal grooves, the rejection edge 148 of the seal is effectively held in particle rejection engagement with the cylindrical sealing surface of the shaft 146. Furthermore, when the seal is deformed in the radial direction, the outer radial corners of the seal 13
9 contacts the support surface 141 of the seal groove to support the seal in the groove to prevent tilting.

As shown in FIGS. 8 and 8A, the housing 150
Defines a seal groove 152, the smallest diameter portion 1 of the tapered or frusto-conical groove surface 154 of this seal groove.
56 corresponds to the outer diameter portion 158 of the dynamic pressure seal 160. 8th
When the seal is subjected to radial compression as shown in the figure,
The radially outer portion of the seal is deformed and abuts against the tapered groove surface 154, and the radially inner dynamic pressure lip (dynamic pressure peripheral seal protrusion) 162 is also deformed. Since the seal is deformed only on the contaminant side and forms a hydrostatic seal with the surface 154, the compression deformation of the seal is substantially symmetrical, and corresponding symmetry is obtained in the inner and outer contact force profiles. Accordingly, the seal does not have a large tendency to tilt or twist within seal groove 152. Since there is no twist, the reject rim 164 maintains optimal engagement with the cylindrical outer surface 166 of the shaft. In addition, the outer corner 151 of the seal is in supporting engagement with the support surface 154 as shown in FIG.

9 and 9A, the dynamic pressure seal 168 of substantially the same structure in FIG. 8 is a general-purpose dynamic pressure rotary shaft seal, and as shown in FIG.
In radial compression. The housing defining the seal groove defines a substantially cylindrical compression portion 174 of the seal that extends laterally a distance substantially equal to the average compression width of the hydrodynamic sealing surface 176. The seal groove has a tapered inner surface portion 178 that cooperates with a radially outer peripheral surface 180 of the seal 168 to form an outer peripheral relief area 182 of the seal. Because the seal 168 is in radial compression between the cylindrical cylindrical sealing surface 184 and the radially opposite compression surface 174, the seal deforms substantially symmetrically, even at the interface contact force profile on both sides. A corresponding symmetry is obtained, and the seal does not have a tendency to be highly inclined or twisted in the seal groove 170. Exclusion edge
The 186 remains scraped and engaged with the cylindrical surface of the shaft, effectively preventing the entry of contaminants into the hydrostatic sealing interface. Radially outer corners 181 are in supporting engagement with support surface 178 to prevent clockwise twisting of the seal in the groove.

When the seal is subjected to a radial compression force, its hydrodynamic sealing edge assumes the shape most effective for lubricant wedge action and contaminant rejection. As shown in FIGS. 10 and 10A,
A rotating shaft 190 extends into the housing structure 188. The housing defines an internal seal groove 192 in which a ring-shaped dynamic pressure seal 194 is located. The uncompressed state of the seal is shown in FIG. 10A, and the radially compressed state of the seal is shown in FIG. In the case of the seal shown in FIG. 10A, the inner surface 195 in the non-compressed state and the dynamic pressure sealing surface 196 are arranged in an angular relationship with the horizontal. The housing defining the seal groove 192 defines a cylindrical radially outer inner surface 198 which moves the radially outer portion of the seal as shown in FIG. At the same time, this radial compression causes the inner periphery of the seal to assume the shape of a shaft as shown in FIG. 10 and the inner periphery of the seal to assume a shape suitable for effective hydrodynamic lubricant wedge action, Also, the circular edge 197 is kept away from the shaft. In this case, it should be noted that the shape of the compressed seal is somewhat twisted, but this twisted shape causes the dynamic pressure motion of the lubricant at the dynamic pressure interface and the dynamic sealing and radiation at the dynamic pressure interface. Suitable for hydrostatic sealing of the outer part in the direction. Furthermore, the exclusion edge of the seal
200 is in efficient scraping engagement with the cylindrical surface of shaft 190. Thus, even when radial compression is applied to the seal, the steep exclusion edge is kept in contaminant exclusion contact with the sealing surface of the shaft because of the internal shape of the seal.

In the embodiment of the invention shown in FIG. 19, the hydraulic seal 201 is substantially identical to the embodiment of FIG. 2A,
The radially outer hydrostatic lips 203 to 204 form a cone. Such a conical lip structure produces a progressively increasing compressive force toward the outer surface of the seal 205, which can withstand the large fluctuations in torsional action resulting from stacking tolerances and lateral offset of the shaft. At the same time, the sealing contact can be maintained at the engagement surface between the packing gland (not shown) and the shaft (not shown). The conical shape from 203 to 204 does not substantially increase the degree of compression between the hydrodynamic sealing lip 206 and the opposing surface of the shaft (not shown) near the hydrodynamic pumping Pumping zeometry 208 to reduce
Does not cause excessive flattening or distortion.

However, the aforementioned conical portion increases the contact pressure between the dynamic pressure lip 206 and the opposing surface of the shaft near the sharp exclusion edge 207, thus eliminating this sharp edge. Improves action. However, as mentioned above, this conical portion does not substantially increase the contact pressure between the dynamic pressure lip and the opposing surface of the shaft, and therefore does not adversely affect the lubricant pumping action. . The advantage of the embodiment shown in FIG. 19 is that the hydrostatic lip
This can also be achieved by providing a conical zeometry between 208 and 207 of the dynamic pressure lip (dynamic pressure seal projection) 206 instead of 2. Such an advantage can also be obtained by applying conical zeometry to both the dynamic pressure lip (dynamic pressure peripheral seal projection) and the static pressure lip (static pressure peripheral seal compressed portion).

Analysis of this embodiment has shown that such a cross section has a high degree of stability within the gland without the tendency of the seal to twist. As with the embodiment of FIG. 2A, the inherent stability of the embodiment shown in FIG. 19 is such that the hydrostatic lip (the hydrostatic circumferential seal projection) compresses the hydrostatic lip (the hydrostatic circumferential seal compressed portion). )). When the seal is compressed, the interfacial contact force profile of both lips and the magnitude and location of the degree of deformation are very similar, so that the seal does not tend to twist too much inside the gland and an equilibrium is obtained.

The seal is deformed so that its hydrodynamic sealing edge is shaped appropriately for efficient lubricant wedge action and contaminant rejection, and restrains the seal from meandering during packing. Is done.

The uncompressed state of the sealing element 209 is shown in FIG. 20A, and the compressed state is shown in FIG.

As seen in FIG. 20, a rotating shaft 211 extends into the housing structure 210. The seal groove 212 defined by the housing has a lubricant sidewall 213 and a contaminant sidewall 214. In this seal groove 212 a dynamic pressure sealing element 209 on the ring is arranged.

In the uncompressed state of FIG. 20A, the hydrodynamic seal zeometry is indicated at 215. Dynamic pressure lip (dynamic pressure peripheral seal projection) 215
The lubricant side defines an axially varying dynamic pressure zeometry 216, and the contaminant side of the dynamic lip defines a sharp dropout edge 217. The hydrodynamic sealing surface 218 of the hydrodynamic lip is arranged in an angular relationship to the axis of the seal and extends from the hydrodynamic zeometry 215 to the steep exclusion zeometry 217.
In the uncompressed state, the inner surface 219 is arranged in an angular relationship with respect to the axis of the seal. The axial end face of the ring-shaped cross section of the seal defines a lubricant side 220 and a contaminant side 221. The contaminant sides are sloped near the rejection edge 217 so that the angle 222 between the dynamic sealing surface and the contaminant sides does not form an acute angle.

In the compressed state shown in FIG. 20A, the housing structure forms a seal groove 212 and a cylindrical radial outer surface 2.
The outer surface 223 moves the radially outer portion of the seal. At the same time, radial compression moves the seal and seals the lubricant side 220 to the seal groove wall 2
13 to bring the contaminant side 221 of the seal into contact with the contaminant side groove wall 214. The contact between the sides of the seal and the wall of the groove constrains the seal and prevents its meandering without applying additional lateral pressure to the seal, thus eliminating the reject edge 217 at the rotational speed of the shaft. It is restrained against warping, thus preventing collisions which would cause wear of the exclusion edge and extending the life of the seal.

Because the hydrodynamic sealing surface 218 is inclined in the uncompressed state, the sharp edge 217 remains in contact with the shaft even when the seal is twisted in the compressed state, and can perform its intended rejection function. Also, the inner portion 219 of the seal is inclined in a non-compressed state and is spaced from the axis in a compressed state despite the twisting action of the seal. In the compressed state, the seal is twisted, but the inclined surface of the dynamic pressure projection cooperates,
Maintaining full contact between the seal and the shaft at the location of the steep exclusion edge 217 maintains effective exclusion.

As is clear from the above. The present invention is a structure suitable for accomplishing all of the above objects and all other objects and features of this type of device.

The present invention is not limited to the above description, but can be arbitrarily changed and implemented within the scope of the gist.

 ────────────────────────────────────────────────── (5) References US Patent 3,970,321 (US, A) US Patent 3,980,309 (US, A) US Patent 4,089,533 (US, A)

Claims (11)

(57) [Claims] 1. A rotary lubricating rotary diaphragm packing type seal for rotatably sealing a sealing surface, removing contaminants and preventing entry into a lubricating area, with a rotatable annular sealing surface between a lubricating side and a contaminated side. In the element: (a) having an inner and outer radial peripheral surface, one of which is suitable for forming a hydrodynamic sealing interface with an annular sealing surface rotatable with respect to this peripheral surface; And the other peripheral surface is suitable for forming a peripheral surface sealing engagement with a hydrostatic seal packing holding surface radially spaced from the rotatable annular sealing surface. And a generally ring-shaped member made of an elastic sealing material having an axial end defining a boundary between the lubricant side and the contaminant side; and (b) the inner and outer radiation of the ring-shaped member. On one side of the circumference The one peripheral surface is adapted to seal with the rotatable annular seal surface and to cause a dynamic pressure wedge effect on the lubricant filled therein in the dynamic pressure seal interface. The lubricant has a shape such that the diameter of the one peripheral surface fluctuates in the axial direction on the lubricant side, and a dynamic edge defining an annular edge for removing contaminants from the hydrodynamic seal interface on the contaminants side. (C) provided on the other of the inner and outer radial peripheral surfaces of the ring-shaped member made of an elastic sealing material, and pressed by the hydrostatic seal packing pressing surface to be compressed. At the time, an annular hydrostatic circumferential seal compressed portion which is made to perform a seal engagement together with the hydrostatic seal packing holding surface and is arranged on a radially opposite side of the hydrodynamic circumferential seal projection,
(D) the dynamic pressure-peripheral seal protrusion and the static pressure-peripheral seal compressed part are disposed at radially symmetric positions with the ring-shaped member interposed therebetween, and the ring-shaped member made of an elastic seal material is provided. Receives the same strain during radial compression between the rotatable annular seal surface and the hydrostatic seal packing holding surface, whereby the ring-shaped member made of an elastic seal material is radially compressed. A dynamic-pressure lubricating rotary throttle packing type sealing element which is not subject to compression-induced torsion when subjected to pressure. 2. The dynamic-pressure lubricating rotary throttle packing type according to claim 1, wherein the compressed portion of the static pressure peripheral seal is formed to protrude from the other peripheral surface in a non-compressed state. Seal element. 3. The hydrodynamic lubricating type rotating device according to claim 1, wherein an outer peripheral surface of said compressed portion of said static pressure peripheral seal is in the same plane as said other peripheral surface in a non-compressed state. Aperture packing type seal element. 4. The hydrostatic lubricating type rotary throttle according to claim 4, wherein said annular hydrostatic seal compressed portion and dynamic pressure peripheral seal projection have equal compression of said hydrostatic peripheral seal compressed portion and dynamic pressure peripheral seal projection. 4. The hydrodynamic lubrication according to claim 2, wherein the ring-shaped member made of a sealing material forming the packing-type sealing element cooperates to apply a compressive load such that twisting does not occur. Rotary diaphragm packing type seal element. 5. The dynamic pressure-peripheral seal projection and the static pressure-peripheral seal compressed portion are both disposed on the contaminant side of the ring-shaped member as a whole and made of an elastic sealing material and disposed on the opposite side in the radial direction. The sealing element according to claim 2 or 3, wherein the seal element is a rotary diaphragm packing type. 6. The dynamic pressure peripheral seal projection and the static pressure peripheral seal compressed portion each form an annular exclusion edge on the contaminant side of the ring-shaped member as a whole, and the annular exclusion edge is a ring. 6. The dynamic pressure lubricating rotary throttle packing type sealing element according to claim 5, wherein the sealing element is arranged on the same plane orthogonal to the axis of the member. 7. A radial compression of said hydrodynamic lubricating rotary throttle packing type sealing element between an inner surface shape of said annular seal packing retainer and said rotatable annular sealing surface. The inner and outer peripheral surfaces having the hydrodynamic peripheral seal projections so that a deformed state of the seal substantially bidirectionally symmetrical at the hydrodynamic peripheral seal protrusion and the hydrostatic peripheral seal compressed portion is generated. One of them retains a cross-sectional shape capable of maintaining a dynamic wedge action of a lubricant in the dynamic seal part and an elimination action of eliminating contaminants from the dynamic seal part,
The dynamic pressure lubricating die according to claim 2 or 3, wherein an annular seal packing retainer that forms an inner surface shape for compression seal engagement with the compressed part of the static pressure peripheral seal is defined. Rotary diaphragm packing type seal element. 8. A seal engaging circumferential surface for radial compression of said annular hydrostatic circumferentially compressed portion in order to achieve radial compression of said generally ring-shaped member without occurrence of compression-induced torsion. Having a housing having an annular seal packing holding surface having
4. A seal element according to claim 2, wherein said seal element is a rotary diaphragm packing. 9. A static arrangement defining a seal including a gland having a hydrostatic seal surface for lubricating the seal surface and preventing contaminants from entering the lubrication area between the lubrication side and the contaminated side. A hydrodynamically lubricated rotary throttle packing type seal element used to seal between the element and a rotating component forming an annular dynamic sealing surface, wherein: (a) inner and outer circumferential surfaces are formed; A first axial end face in contact with contaminants and a second end face in contact with the lubricant;
A ring-shaped member made of an elastic sealing material having an axial end face formed thereon; and (b) a first radially projecting direction projecting from the generally annular ring-shaped member made of a sealing material. A first axial direction for contacting the contaminant and making rubbing contact with the annular dynamic pressure sealing surface of the rotating component; An extruded edge is formed at the end of the annular dynamic pressure seal surface, and a wedge-shaped lubricant is inserted into an interface between the dynamic pressure seal protrusion and the annular dynamic pressure seal surface of the rotating shaft. (C) at least one hydrostatic seal projection having an axially varying tapered shoulder for making dynamic pressure wedge contact with the lubricant; Seal contact A first radial projection extending from the entire annular ring member made of a sealing material and disposed in a radially opposed relationship with the dynamic pressure seal projection; and the dynamic pressure seal surface and the static pressure seal surface When subjected to radial compression between the hydrodynamic projections, the hydrodynamic projections undergo radial deformation that is substantially equal to the dynamic pressure projections, and the overall ring-shaped member made of an elastic sealing material is used for the dynamic pressure sealing element and the static pressure sealing surface. Wherein the dynamic pressure sealing element is provided with at least one annular hydrostatic seal projection which is free from torsion due to radial deformation by being subjected to radial compression between and. Dynamic pressure lubrication type rotary diaphragm packing type sealing element. 10. The dynamic and static pressure outer peripheral seal projections
Both have an annular shape and are disposed on the contaminant side of the overall ring-shaped member made of a sealing material,
When the dynamic pressure seal projection is radially deformed by compression of the rotating shaft by the dynamic pressure seal surface,
10. The dynamic pressure lubricating rotary throttle packing type seal according to claim 9, wherein the dynamic pressure seal projection forms an axially changing tapered shoulder for forming wedge contact with a lubricant. element. 11. The annular component of the generally ring-shaped member of sealing material such that radial compression does not cause twisting of the generally ring-shaped member of sealing material. 10. A packing gland which forms an inner circumferential annular sealing surface containing a seal for radial compression in cooperation with the static pressure seal projection and the static component. The described dynamic pressure lubricating rotary throttle packing type sealing element.