WO2009032248A1 - Joint d'étanchéité rotatif avec amélioration de la distribution de film - Google Patents

Joint d'étanchéité rotatif avec amélioration de la distribution de film Download PDF

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Publication number
WO2009032248A1
WO2009032248A1 PCT/US2008/010320 US2008010320W WO2009032248A1 WO 2009032248 A1 WO2009032248 A1 WO 2009032248A1 US 2008010320 W US2008010320 W US 2008010320W WO 2009032248 A1 WO2009032248 A1 WO 2009032248A1
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WIPO (PCT)
Prior art keywords
seal
lip
dynamic
locations
footprint
Prior art date
Application number
PCT/US2008/010320
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English (en)
Inventor
Lannie L. Dietle
John E. Schroeder
Original Assignee
Kalsi Engineering, Inc.
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
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Publication date
Application filed by Kalsi Engineering, Inc. filed Critical Kalsi Engineering, Inc.
Priority to CA2697678A priority Critical patent/CA2697678C/fr
Publication of WO2009032248A1 publication Critical patent/WO2009032248A1/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/32—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings
    • F16J15/3244—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings with hydrodynamic pumping action
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/164—Sealings between relatively-moving surfaces the sealing action depending on movements; pressure difference, temperature or presence of leaking fluid
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/32—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings
    • F16J15/3204—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings with at least one lip
    • F16J15/3216—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings with at least one lip supported in a direction parallel to the surfaces
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/32—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings
    • F16J15/324—Arrangements for lubrication or cooling of the sealing itself
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/16—Sealings between relatively-moving surfaces
    • F16J15/32—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings
    • F16J15/3284—Sealings between relatively-moving surfaces with elastic sealings, e.g. O-rings characterised by their structure; Selection of materials

Definitions

  • This invention relates to hydrodynamic rotary seals for bi-directional rotation that are used to retain a lubricant and exclude an environment. More specifically, this invention relates to cooperative features that improve seal lubrication in conditions such as high operating temperature, skew-resisting confinement, high differential pressure, high initial compression,
  • Page l of 56 adverse tolerance accumulation, circumferential compression, high modulus seal materials, thin viscosity lubricants, third body seal surface wear, and/or material swell (collectively referred to as “severe operating conditions").
  • hydrodynamic seals Upon installation in a compressed condition, hydrodynamic seals define a "footprint" representing the shape of the "dynamic sealing interface,” and the two terms are generally interchangeable. Examples of footprints are shown in FIG. 2 of assignee's U.S. Pat. 4,610,319 and FIG. 13 of assignee's U.S. Pat. 5,230,520.
  • skew-induced wear mechanism described and illustrated in FIG. 3-27 of the Kalsi Seals Handbook, Rev. 1 is addressed with skew-resisting confinement of the seal, which increases interfacial contact pressure and footprint spread.
  • the term "skew-resisting confinement,” as used herein, encompasses (1) constraint imposed by seal contact with fixed location gland walls as disclosed in U.S. Pat. 6,315,302, and (2) spring-loading through a moveable gland wall, as disclosed in FIG. 3-28 of the Kalsi Seals Handbook, Rev. 1.
  • FIG. 13 uses a curved arrow to illustrate the conventional wisdom that a normal velocity component VN urges the lubricant toward the environment.
  • VN has caused undue focus on inlet efficiency over the years, and diverted attention from finding other potential lubrication factors.
  • Such conventional wisdom of how these seals operate has been repeated in numerous other patents and commercial literature. See, for example, U.S. Patents 5,678,829 (col. 4, lines 14-33), 5,738,358 (col.
  • U.S. Patent 6,109,618 teaches the use of abrupt trailing edge geometries, that are unsuitable as hydrodynamic inlets, on seals suitable only for uni-directional rotation. This abrupt geometry is on the trailing edges of the waves, and is coupled with a very gently converging inlet geometry on the leading edges. Due to the high hydrodynamic leakage of such geometry, and the small reservoir size of downhole tools, downhole seals cannot employ such geometries.
  • the prior art seals are constructed from elastomers which suffer accelerated degradation at elevated temperature. For example, media resistance problems, gas permeation, swelling, compression set, and pressure related extrusion damage all become worse at elevated temperature.
  • a bi-directional rotation seal that operates with less torque and produces less seal- generated heat would be desirable, in order to moderate such degradation.
  • Circumferential slippage of a seal with respect to its groove occurs more often with large diameter seals because the moment arms between the static and dynamic sealing interfaces are more nearly equal, and the static sealing interface has less mechanical advantage.
  • Rotational slippage is particularly undesirable in large diameter seals because the slippage can vary around the circumference of the seal, causing undesirable localized stretching. It is also undesirable in seals exposed to high differential pressure because slippage can accelerate seal extrusion damage. Slippage is exacerbated by seal or shaft wear because such wear increases running torque. A bi-directional seal that has lower running torque and more resistance to wear is therefore desirable.
  • the present invention relates to generally circular rotary seals that are suitable for bidirectional rotation, and overcome the aforementioned prior art problems.
  • the seals are used to establish sealing between a machine component (such as a housing) and a relatively rotatable surface (such as a shaft), in order to separate a lubricating media from an environment.
  • Seal geometry on a dynamic lip interacts with the lubricating media during relative rotation to wedge a lubricating film into the dynamic sealing interface between the seal and the relatively rotatable surface.
  • a portion of the lubricating film migrates toward, and into the environment and thus provides a contaminant flushing action.
  • the rotary seal includes a dynamic lip that deforms when compressed into sealing engagement with the relatively rotatable surface, defining a hydrodynamic wedging angle with respect to the relatively rotatable surface, and defining an interfacial contact footprint of generally circular configuration but varying in width.
  • a non-circular (i.e., wavy) footprint edge hydrodynamically wedges the lubricating film into the interfacial contact footprint.
  • EWH- EWH a newly established variable referred to as EWH- EWH is used herein for a dimension that represents the
  • Dimension B 2 is the dimension from a second footprint edge to a location P 2 defining the maximum interfacial contact pressure at the widest footprint location and Width Wj is the footprint width at the narrowest footprint location.
  • the value of E W H is calculated as Dimension B 2 minus the Width Wj, and the result may be a positive or negative number, depending on the circumstances.
  • Another embodiment of the present invention is a generally circular, hydrodynamically lubricating rotary seal that accomplishes improved lubrication through the cooperative benefits of a modified zig-zag wave pattern, a variably sized inlet curvature that is a tighter curvature near the widest parts of the dynamic lip, and a less tight curvature near the narrower portions of the dynamic lip, and a dynamic lip flank that is more steeply sloped near the widest parts of the dynamic lip, and less steep near the narrower parts of the dynamic lip.
  • the modified zig-zag wave pattern improves interfacial contact pressure gradients and the orientation and/or location of the pressure gradients in critical locations.
  • variable radius controls the magnitude of the pressure at a critical location, and provides improved inlet convergence, while enhancing factors that contribute to lubrication in severe operating conditions.
  • variable slope of the dynamic lip flank provides a number of benefits related to a variety of seal performance issues, the most important of which is to minimize seal volume for improved compatibility with skew-resisting confinement.
  • the seal preferably provides a dynamic exclusionary intersection of abrupt substantially circular form that provides the interfacial contact footprint with an environment edge that resists environmental intrusion.
  • the seal can be configured for dynamic sealing against a shaft, a bore, or a face. Simplified embodiments are possible wherein one or more features of the preferred embodiment are omitted.
  • One objective of this invention is to provide a hydrodynamic rotary seal having low torque for reduced wear and heat generation. Another objective of an embodiment is improved distribution of lubricant across the dynamic sealing interface, particularly at high operating temperatures. Another objective of yet another embodiment is to conserve void volume within the seal gland, to provide adequate room for seal thermal expansion, considering seal tolerances and as-manufactured variations in the coefficient of thermal expansion of the sealing material, with a view toward improved accommodation of skew-resisting confinement.
  • FIG. 1 is a graph schematically representing a typical interfacial contact pressure plot at a circumferential location of a hydrodynamic seal
  • FIG. 2 is a fragmentary cross-sectional view of a ring-shaped hydrodynamic seal according to a preferred embodiment of the present invention, the seal shown in an installed, compressed condition;
  • FIG. 2A is a fragmentary view of the hydrodynamic seal of FIG. 2 in an uncompressed condition
  • FIG. 2B is a section view taken along line 2B-2B of FIGS. 2A and 5;
  • FIG. 2C is a section view taken along line 2C-2C of FIGS. 2A, 3, 4, 5, 6, 7 and 8;
  • FIG. 2D is a section view taken along line 2D-2D of FIGS. 2A, 3, 4, 5, 6 and 7;
  • FIG. 2E is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 2;
  • FIG. 3 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 3A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 3;
  • FIG. 4 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 4A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 4;
  • FIG. 4B is a section view taken along line 4B-4B of FIGS. 4, 7 and 8;
  • FIG. 5 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 5A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 5;
  • FIG. 6 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 6A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 6;
  • FIG. 7 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 7A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 7;
  • FIG. 7B is a graph representing the interfacial contact pressures at selected circumferential slices of the hydrodynamic rotary seal of FIGS. 2A and 7;
  • FIG. 8 is a fragmentary view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal in an uncompressed condition;
  • FIG. 8A is a fragmentary illustration of the interfacial contact footprint of the hydrodynamic seal of FIG. 8;
  • FIG. 9 is a fragmentary cross-sectional view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal shown in an installed, compressed condition;
  • FIG. 9A is a fragmentary view of the hydrodynamic seal of FIG. 9 in an uncompressed condition
  • FIG. 10 is a fragmentary shaded perspective view of an alternative embodiment of the hydrodynamic seal showing a hydrodynamic inlet portion of a wave.
  • FIG. 11 is a fragmentary cross-sectional view of a hydrodynamic seal according to another preferred embodiment of the present invention, the seal shown in an installed, compressed condition.
  • FIGURE 1 is a graph that schematically represents an interfacial contact pressure plot at any circumferential location of a typical seal, for example, manufactured according to one of assignee's U.S. Patents 4,610,319, 5,230,520, 6,315,302, 6,382,634, and so forth.
  • the proportions of a contact pressure plot will vary depending on specific seal geometry and analysis constraints, but the general plot characteristics are captured in FIG. 1.
  • the plot has a first footprint edge L and second footprint edge E, which correspond to the lubricant-side and environment-side edges, respectively, of the dynamic sealing interface/footprint.
  • the direction of relative rotation between the seal and the mating relatively rotatable surface is normal (perpendicular) to the axis labeled "interfacial width", and normal (perpendicular) to the page on which the figure is printed.
  • the inlet contact pressure rises from zero at the first footprint edge L to a maximum at Location P.
  • Location P is remote from the first footprint edge L by Dimension A, and is remote from the second footprint edge E by Dimension B.
  • Width W represents the local width of the dynamic sealing interface/footprint.
  • the footprint edge represented by second footprint edge E is substantially circular.
  • Width W varies about the circumference of the seal from a narrowest location defined by Width W) to a widest location defined by Width W 2 .
  • the variation of Width W has been sinusoidal.
  • EWH is a newly established variable that is used herein for a dimension that represents the difference in size between Dimension B 2 and Width Wj. In other words, the value of EWH is calculated as Dimension B 2 minus the Width W], and the result may be a positive or negative number, depending on the circumstances.
  • the lubricant film thickness is uneven across the Width W of the dynamic interface, and surface asperity contact sometimes occurs. For example, a portion of the footprint that is circumferentially aligned with Width Wj suffers film disruption due to the aforementioned circumferential contact pressure zone orientation, the sheer magnitude of contact pressure near Location Pi, and due to unfavorable contact pressure gradients.
  • FIGURE 5 of U.S. Pat. 4,610,319 shows a wave pattern that addresses the aforementioned generally circumferential contact pressure zone orientation, but it does not address the zone of elevated contact pressure near Location P]. If the teachings of FIG. 5 of the
  • FIGURES 2A, 2B, 2C, 2F and 9 of U.S. Pat. 6,109,618 teach the use of a similar wave for uni-directional rotation. These figures fail to address lubricating problems in the vicinity of Location ? ⁇ .
  • the narrowest part of the dynamic lip is dominated by an abrupt restrictive diverter 250.
  • the abrupt restrictive diverter 250 is in the form of a corner/facet between dynamic sealing surface 226 and wavy lubricant-side 230 of circular dynamic sealing lip 224, and in FIG.
  • the abrupt restrictive diverter 250 is in the form of a sharp projection.
  • the abrupt restrictive diverter 250 causes contact pressure to skyrocket at the narrowest parts of the dynamic lip, making problems worse.
  • the salient issues were clearly missed because the contact pressure zone 262 in FIG. 2F does not extend to Location Pj, and the illustrated edge of the zone is circumferentially oriented at the narrowest points of the footprint.
  • the footprint edge radii at the narrowest parts of the footprint are large— indicating that any corresponding seal fillets are much larger than needed to eliminate the surface facets inherent to FIG. 5 of U.S. Pat. 4,610,319.
  • U.S. Patents 6,109,618 and 6,685,194 teach different ways of implementing a variable inlet curvature on the leading edge of a uni-directional hydrodynamic wave. Both patents teach the use of a curvature that is most abrupt at the narrowest parts of the dynamic lip (for example, see FIGS. 2 to 2E of the '618 patent, and FIGS. 4 to 4C, 5A, 7, 8, 8A, 9 and 9A of the '194 patent). Because the inlet is more abrupt at the narrowest parts of the dynamic lip, it exacerbates contact pressure issues near Location Pj.
  • HSN Highly saturated nitrile
  • seals made and employed in accordance with U.S. Pat. 6,315,302 suffer from under-lubrication at higher temperatures that are within the generally understood operating temperature limits of the elastomer (the 25O 0 F temperature stated in the '302 patent at col. 11, line 65 through col. 12, line 2 is well within the generally understood operating limits of the elastomer).
  • TFE/P-FKM tetrafluoroethylene and propylene copolymer- flurocarbon rubber
  • Elastomers have a high coefficient of thermal expansion. Because there is more material at the widest parts of the dynamic lip, part of the differential thermal expansion between the seal and the housing is relieved circumferentially, causing material displacement from the widest to the narrowest parts of the dynamic lip, thereby increasing Width Wi relative to Dimension B 2 .
  • the term "un-swept zone” refers to that portion of the footprint that is circumferentially aligned with Width Wi, and the term “swept zone” refers to all the other area of the footprint. In other words the swept zone is that portion of the footprint that is circumferentially aligned with the footprint wave height.
  • the swept zone is directly lubricated by the sweep of the First Footprint Edge L across the lubricant- wetted shaft.
  • FIGURE 13 of U.S. Pat. 6,109,618 shows how pervasive the conventional wisdom concerning a gentle inlet convergence has been.
  • a seal intended for bi-directional rotation is shown where the inlet geometry produces extremely effective in-pumping by virtue of gentle convergence with the shaft, but unfortunately another gentle convergence allows the lubricant film to escape at the trailing edge of the wave.
  • FIGURE 13 also demonstrates a lack of understanding of the problems that circumferentially-oriented contact zones cause. As shown by the dashed line representation of the tangency location, the seal of FIG. 13 has circumferentially-oriented zones of contact pressure that extend over most of the circumference of the seal.
  • FIGS. 13A and 13B show that the contact pressure within these zones are relatively high because of the abrupt nature of the lip flank curvature, as shown in the section views of FIGS. 13A and 13B.
  • 6,685,194 also demonstrate how pervasive the conventional wisdom has been, in relation to achieving a gentle inlet convergence that was believed, in conjunction with VN, to create an ideal situation for lubricating the interfacial contact footprint. Such geometries were eventually found to suffer from under- lubrication in severe service conditions. After years of research, the cause has recently been determined to be the result of the value of EWH approaching or equaling zero, or becoming negative.
  • the lubrication is due to secondary factors such as side leakage from the swept zone (related to the fact that film thickness tends to decay gradually, rather than abruptly, due to the relative stiffness of the seal material; see Abstract, U.S. Pat. 6,109,618), and such as macro-lubrication from surface finish affects.
  • each molded wave is substantially identical, which means that all instances of Dimension Bj are substantially identical.
  • the various aforementioned factors cooperate to thin the film in the un-swept zone.
  • the resulting seal-generated heat exacerbates the aforementioned increase in the size of Width W 1.
  • U.S. Pat. 6,315,302 entitled “Skew Resisting Hydrodynamic Seal,” teaches conservation of void volume to accommodate skew-resisting confinement. From a void volume conservation standpoint, it is desirable to avoid the condition shown in FIGS. 2A and 6 of U.S. Pat. 6,382,634 and FIG. 11 of U.S. Pat. 5,230,520, where the lubricant side flank is truncated by the lubricant end of the seal at the widest part of the dynamic lip, leaving very little void volume near that location. Likewise, it is desirable to avoid the condition shown in FIGS. 4-4C of U.S. Pat.
  • the initial compression also causes circumferential compression, which is increased by thermal expansion. Since the seal circumference is relatively long compared to the seal cross-section, circumferential compression can cause buckling in a manner similar to the classic textbook example of a long, slender structural column under compressive loading. This buckling tendency is augmented by the variable stiffness of the prior art seal about its circumference that is caused by the varying dynamic lip width.
  • FIGURES 2-2E represent a preferred embodiment of the present invention. These figures should be studied together to best understand the preferred embodiment. Features throughout this specification that are represented by like numbers have the same function.
  • FIGURE 2 is a fragmentary cross-sectional view that provides a general overview of how a preferred embodiment of the present invention may be employed when assembled into a machine that is shown generally at 2.
  • the machine 2 includes a first machine component 4 and a second machine component 6 that defines a relatively rotatable surface 8.
  • the first machine component 4 and the second machine component 6 together typically define at least a portion of a chamber for locating a first fluid 12.
  • a rotary seal shown generally at 10, establishes sealing engagement with the relatively rotatable surface 8, to retain the first fluid 12, to partition the first fluid 12 from a second fluid 14, and typically to exclude the second fluid 14.
  • the term "fluid" has its broadest meaning, encompassing both liquids and gases.
  • the first fluid 12 is preferably a liquid-type lubricant such as a synthetic or natural oil, although other fluids are also suitable in some applications.
  • the second fluid 14 may be any type of environment that the rotary seal 10 may be exposed to in service, such as any type of liquid or gaseous environment including, but not limited to, a lubricating media, a process media, a drilling fluid, an atmosphere, seawater, a partial vacuum, etc.
  • the rotary seal 10 is of generally circular, ring-like configuration and includes at least one dynamic Hp 16 that is also generally circular in form, and is disposed in facing relation to the relatively rotatable surface 8.
  • the cutting plane of the cross-section is aligned with and passes through the theoretical axis/centerline of the rotary seal 10; i.e., the theoretical axis lies on the cutting plane.
  • the dynamic lip 16 When exclusion of the second fluid 14 is desired, the dynamic lip 16 preferably incorporates a dynamic exclusionary intersection 44 (sometimes called the "exclusion edge") of abrupt substantially circular form that is substantially aligned with the direction of relative rotation, and is adapted to exclude the second fluid 14, as taught by U.S. Pat. 4,610,319.
  • a dynamic exclusionary intersection 44 (sometimes called the "exclusion edge") of abrupt substantially circular form that is substantially aligned with the direction of relative rotation, and is adapted to exclude the second fluid 14, as taught by U.S. Pat. 4,610,319.
  • the dynamic exclusionary intersection 44 develops substantially no hydrodynamic wedging activity during relative rotation, and presents a scraping edge to exclude the second fluid 14 in the event of relative motion that is perpendicular to the direction of relative rotation.
  • the rotary seals of the present invention may incorporate one or more seal materials or components without departing from the spirit or scope of the invention, and may be composed of any suitable sealing material or materials, including plastics and elastomeric or rubber-like materials including, but not limited to, carbon, fiber or fabric reinforced elastomers.
  • the rotary seals may be of monolithic integral, one piece construction as shown in FIGS. 9, 9A and 10, or may also incorporate different materials bonded, inter-fitted, co-vulcanized or otherwise joined together to form a composite structure, such as shown in U.S. Patents 5,738,358, 6,315,302, 6,685,194, 6,767,016 and U.S. Pat. Appl. Publications 2006/0214379 and 2006/0214380.
  • at least part of the seal is constructed of a resilient material, such as an elastomer.
  • Elastomers used in seal construction are compounds that include base elastomers such as, but not limited to, HNBR (hydrogenated nitrile, also known as HSN), FKM (fluorocarbon rubber), TFE/P (also known as FEPM) and EPDM.
  • the elastomers may include other compounding ingredients such as, but not limited to, fillers, lubricants, processing aids, anti- degradants, vulcanizing agents, accelerators and activators. The effects of the ingredients are generally understood by individuals having ordinary skill in the art of compounding elastomers.
  • the invention can, if desired, incorporate an energizer to load the dynamic lip 16 against the relatively rotatable surface 8.
  • the energizer can take any of a number of suitable forms known in the art including, but not limited to, elastomeric rings and various forms of springs such as garter springs, canted coil springs, and cantilever springs, without departing from the scope or spirit of the invention.
  • the energizer can be located by or within an annular recess of any suitable form. For examples of such energizers and recesses, see U.S. Patents 6,685,194 and 7,052,020, and U.S. Pat. Appl. Publications 2006/0214380 and 2007/0013143.
  • FIGURES 2, 2B, 2C, 2D, 4B and 11 portray seal cross-sections having the high temperature, composite construction taught in U.S. Pat. Appl. Pub. 2006/0214379, where preferably a first material layer 48 of TFE/P is co-vulcanized to a second material layer 49 of FKM.
  • the rotary seal 10 is oriented (i.e., positioned) at least in part by the first machine component 4.
  • the machine 2 has a generally circular seal groove that includes a first groove wall 18 and a second groove wall 20 that are in generally opposed relation to one another, and a peripheral groove wall 22.
  • the first groove wall 18 and the second groove wall 20 are often referred to as the "lubricant-side gland wall,” and the "environment-side gland wall,” respectively.
  • the peripheral groove wall 22 can be substantially parallel to the relatively rotatable surface 8 as shown, or all or part of it could be skewed with respect to the relatively rotatable surface 8 as shown, for example, by the prior art of FIGS. 4, 6, 7, 8 or 9 of U.S. Pat. 5,230,520.
  • the seal "groove” is the annular void that is defined by the first groove wall 18, the peripheral groove wall 22, and the second groove wall 20, and the seal “gland” is the generally enclosed annular cavity having a boundary that is defined by the groove and the relatively rotatable surface 8.
  • the peripheral groove wall 22 is located in spaced relation to the relatively rotatable surface 8, and it (or an energizer) compresses at least a portion of the dynamic lip 16 against the relatively rotatable surface 8. This compression establishes an interfacial contact footprint, shown generally at 38, between the dynamic lip 16 and the relatively rotatable surface 8.
  • the footprint 38 has a first footprint edge located generally at L and facing the first fluid 12, and has a second footprint edge located generally at E and facing the second fluid 14.
  • the second footprint edge E is established by compression of the dynamic exclusionary intersection 44 against relatively rotatable surface 8.
  • the compression causes contact pressure at the interface (footprint 38) between the dynamic lip 16 and the relatively rotatable surface 8. Sealing is also established at the interface between a static sealing surface 46 of rotary seal 10 and the peripheral groove wall 22.
  • the contact pressure at the footprint 38 establishes sealing in the same manner as any conventional resilient seal, such as an O-ring or a seal having a lip that is loaded by an energizer.
  • the interfacial contact pressure is related to the degree of compression, the modulus of elasticity of the seal material, and the shape of the rotary seal 10.
  • the first footprint edge L is preferably wavy.
  • Each wave of the footprint 38 has a leading edge and a trailing edge, relative to the direction of relative rotation.
  • the application of the leading edge/trailing edge appellations also reverses.
  • the leading edges of the waves are sites of hydrodynamic wedging action during relative rotation between the dynamic lip 16 and the relatively rotatable surface 8.
  • This hydrodynamic wedging action forces a film of lubricating fluid (i.e., a film of the first fluid 12) into the interfacial contact footprint 38 for lubrication purposes.
  • the dynamic lip 16 slips or hydroplanes on a film of the first fluid 12 during periods of relative rotation.
  • the hydroplaning activity reduces wear and seal-generated heat, and causes a minute flow of the first fluid 12 past the second footprint edge E and into the second fluid 14.
  • the hydroplaning activity stops, and a static sealing relationship is re-established.
  • the second footprint edge E (sometimes called the "environment edge”) is preferably substantially circular and substantially aligned with the possible directions of relative rotation between the dynamic lip 16 and the relatively rotatable surface 8, and does not produce a hydrodynamic wedging action in response to relative rotation between the dynamic lip 16 and the relatively rotatable surface 8.
  • the rotary seal 10 is preferably held in skew-resisting confinement by virtue of simultaneously contacting the first groove wall 18 and the second groove wall 20.
  • the first groove wall 18 is shown in FIG. 2 as a face of a spring-loaded seal loading ring 24 of the general type taught by FIG. 3-28 of the Kalsi Seals Handbook, Rev. 1.
  • the first groove wall 18 is loaded against the rotary seal 10 by a spring 28 that acts on the seal loading ring 24.
  • the spring load is reacted to a retainer 30 of any suitable form.
  • the spring-loading arrangement can take any of a number of suitable forms without departing from the spirit or scope of the invention. For example, a disk or coil spring arrangement could be substituted for the wave spring arrangement.
  • first groove wall 18 and the second groove wall 20 are shown to be in movable relation to one another in FIG. 2, such is not intended to limit the scope of the invention, for the invention admits to other equally suitable arrangements.
  • first groove wall 18 and the second groove wall 20 could be fixed in position relative to each other.
  • the rotary seal 10 is shown as having a contacting relationship with the first groove wall 18 and the second groove wall 20 in FIGS. 2 and 9, the features of the present invention are also advantageous for applications where the rotary seal 10 only contacts one groove wall at a time, as represented by FIG. 11.
  • the rotary seal 10 preferably defines a first seal end 34 that generally faces the first groove wall 18 and first fluid 12, and preferably also defines a second seal end 36 that generally faces the second groove wall 20 and the second fluid 14.
  • the first seal end 34 and the second seal end 36 are often referred to as the "lubricant end” and the "environment end,” respectively, and are preferably in generally opposed relation.
  • the first seal end 34 and second seal end 36 can take other forms without departing from the spirit or scope of the invention.
  • the first seal end 34 could be angulated as taught in U.S. Pat. 6,315,302 at col. 14, lines 22-22, or could be wavy as taught in U.S. Pat. Appl. Pub. 2007/0205563.
  • the second seal end 36 could include a recess for incorporating an energizer, as is well-known in the art.
  • the preferred embodiment of the present invention has application where the first machine component 4, the second machine component 6, or both, are individually rotatable.
  • the direction of relative rotation is normal (perpendicular) to the plane of the cross-section, and approximately concentric to the dynamic exclusionary intersection 44.
  • the theoretical axis of the rotary seal 10 generally coincides with the axis of relative rotation.
  • the relatively rotatable surface 8 In dynamic operation, has relative rotation with respect to dynamic lip 16 and first machine component 4. The relatively rotatable surface 8 slips with respect to dynamic lip 16, causing the interfacial contact footprint 38 to become a dynamic sealing interface. In the absence of relative rotation, the interfacial contact footprint 38 is a static sealing interface.
  • the rotary seal 10 preferably remains stationary relative to the first machine component 4, however, the prior art teaches that hydrodynamic seal embodiments are possible where relative rotation with the first machine component 4 is allowable; for example, see FIGS. 8 and 8A of U.S. Pat. 6,685,194.
  • the rotary seal 10 is shown located in a position that would occur if the pressure of the first fluid 12 were greater than or equal to the pressure of the second fluid 14. In such conditions, the force of the spring 28, and any differential pressure that may be present, forces the rotary seal 10 against the second groove wall 20. Owing to the complimentary shapes of the second seal end 36 and the second groove wall 20, the rotary seal 10 is well supported at all locations except the small clearance gap 52 (often called the "extrusion gap") that exists between the first machine component 4 and the relatively rotatable surface 8.
  • the relatively rotatable surface 8 can take the form of an externally- or internally- oriented substantially cylindrical surface, as desired, with the rotary seal 10 compressed radially between the peripheral groove wall 22 and the relatively rotatable surface 8, in which case the axis of relative rotation would be substantially parallel to relatively rotatable surface 8.
  • the dynamic lip 16 is oriented for compression in a substantially radial direction, and the peripheral groove wall 22 may be of substantially cylindrical configuration, and first groove wall 18, second groove wall 20, first seal end 34 and second seal end 36 may, if desired, be of substantially planar configuration.
  • the dynamic lip 16 is located either on the inner or the outer periphery of the seal, depending on whether the relatively rotatable surface 8 is an external or internal cylindrical surface.
  • the relatively rotatable surface 8 can take the form of a substantially planar surface, with the rotary seal 10 compressed axially between the peripheral groove wall 22 and the relatively rotatable surface 8 in a "face-sealing" arrangement, in which case the axis or relative rotation would be substantially perpendicular to the relatively rotatable surface 8.
  • the dynamic lip 16 would be oriented for compression in a substantially axial direction
  • the peripheral groove wall 22 may be of substantially planar configuration
  • the first groove wall 18, second groove wall 20, first seal end 34 and second seal end 36 may, if desired, be of substantially cylindrical configuration.
  • the hydrodynamic features can be oriented to pump in a radially outward direction if the first fluid 12 is located inward of the dynamic lip 16, or can be oriented to pump in a radially inward direction if the first fluid 12 is located outward of the dynamic lip 16.
  • the backup ring 24 is preferably segmented or split. If split, the ring itself can, if desired, provide the spring force.
  • FIGURE 2A is a fragmentary view of the rotary seal 10 in the uncompressed condition. To minimize curvature-related foreshortening in the illustrations, for ease of understanding, FIGS.
  • FIGURES 2B-2D are section views representative of cutting planes 2B-2B, 2C-2C and 2D-2D, respectively, and represent the uncompressed cross-sectional shape of the rotary seal 10 of the preferred embodiment. These cutting planes are used on FIGS. 2A, 3, 4, 5, 6, 7 and 8.
  • FIGURES 2B, 2D and 4B are section views that are representative of the narrower portions of the dynamic Hp 16, and FIG. 2C is representative of the wider portions of the dynamic lip 16.
  • FIGS. 2B-2D various previously defined portions of rotary seal 10 are labeled for orientation purposes, such as the dynamic lip 16, first seal end 34, second seal end 36, dynamic exclusionary intersection 44, static sealing surface 46, first material layer 48, and second material layer 49.
  • the static sealing surface 46 may, if desired, be defined by a projecting static lip 54 to provide a degree of twist-inhibiting compressive symmetry, as taught by U.S. Pat. 5,230,520.
  • This arrangement provides a recessed surface 55 that, by being recessed, helps to minimize the volume of rotary seal 10, and thereby helps to maximize void volume within the gland.
  • the projecting static lip 54 is preferably oriented in generally opposed relation to the dynamic lip 16. If desired, the embodiments illustrated herein can be simplified by eliminating the projecting static lip 54, such that the static sealing surface 46 is defined by the seal body, as taught by U.S. Pat. 4,610,319.
  • the dynamic lip 16 defines a dynamic surface 56 and a lubricant side flank 58 that are blended by an inlet curvature 60.
  • the dynamic surface 56 and the recessed surface 55 need not be parallel.
  • the lubricant side flank 58 is located in spaced relation with respect to the dynamic exclusionary intersection 44 and the second seal end 36.
  • the inlet curvature 60 can be any suitable curved shape, such as, but not limited to, a radius, a portion of an ellipse, a portion of a sine wave curve, a portion of a parabolic curve, a portion of a cycloid curve, a portion of witch/versiera curves, or combinations thereof.
  • the static sealing surface 46 and/or the dynamic surface 56 can be of tapered configuration as taught by U.S. Pat. 6,767,016.
  • the dynamic surface 56 and the lubricant side flank 58 would intersect at a theoretical intersection 62 that is positioned from the dynamic exclusionary intersection 44 by a distance that varies along the circumference of the rotary seal.
  • the theoretical intersection 62 is non-circular and wavy.
  • the lubricant side flank 58 is also non-circular and wavy, hi keeping with American drafting third angle projection conventional representation, the theoretical intersection 62 is represented by an object line in FIGS. 2A, 3, 4, 5, 6, 7 and 8, even though blended by the inlet curvature 60.
  • this blended intersection illustration convention see paragraph 7.36 and FIG. 7.44(b) on page 213 of the classic drafting textbook "Technical Drawing," 10th edition (Prentice-Hall, Upper Saddle River, N.J.: 1997).
  • the extent of the inlet curvature 60 is represented by a first extent line 64 and a second extent line 66 in FIGS. 2A, 3, 4, 5, 6, 7 and 8.
  • a substantial tangency exists between the inlet curvature 60 and the lubricant side flank 58 at the first extent line 64, and between the inlet curvature 60 and the dynamic surface 56 at the second extent line 66.
  • the first extent line 64 and the second extent line 66 are skewed with respect to the possible directions of relative rotation; in other words, they are non-circular and wavy.
  • second extent line 66 reverses direction at a First Reversing Location R ⁇ and at a
  • Second Reversing Location R 2 and that these locations are preferably staggered by Offset
  • Offset Dimension T Another way to say this is that some of the blending curves 84 are offset with respect to other of the blending curves 84 by Offset Dimension T.
  • the Offset Dimension T and the local cross-sectional geometry of the inlet curvature 60 in FIGS. 2A, 4, 5, 7 and 8 govern the size of the Offset Dimension X that is shown in FIGS. 2E, 4A, 5A, 7A, and 8A, respectively.
  • FIGURE 2E illustrates a fragmentary portion of the footprint (shown generally at 38) of the rotary seal 10 that is portrayed in FIGS. 2-2D.
  • FIGURES 4A, 5 A, 7A and 8 A illustrate a fragmentary portion of the footprint 38 of the rotary seal 10 that is portrayed in FIGS. 4, 5, 7 and 8, respectively.
  • FIGURES 2E, 4A, 5A, 7 A and 8 A use the same nomenclature as FIG. 1, however the subscript "1" has, when necessary, been modified by the addition of an "a” or "b” to designate specific locations of the footprint 38.
  • the dynamic lip 16 has wider lip locations 80 at cutting plane 2C-2C and narrower lip locations 82 at cutting planes 2B-2B and 2D-2D.
  • the inlet curvature 60 preferably varies in curvature about the circumference of the rotary seal 10. At or near the wider lip locations 80, the inlet curvature 60 is preferably a tighter curve, compared to the curve at the narrower lip locations 82.
  • the radius might be smallest in size at and/or near cutting planes 2C-2C (i.e., at and/or near the wider lip locations 80), medium in size at and/or near cutting planes 2B-2B, and largest in size at and/or near cutting plane 2D-2D (i.e., at the narrower lip locations 82).
  • the inlet curvature 60 might be a portion of an ellipse, wherein the major axis varies from being smallest at and/or near the wider lip locations 80 (i.e., at and/or near cutting plane 2C-2C), to being largest at and/or near some of the narrower lip locations 82 (such as cutting plane 2D-2D) while varying to a medium size at and/or near other of the narrower lip locations 82 (such as at cutting planes 2B-2B).
  • the minor and major axes can be identical to each other at and/or near the wider lip locations 80. It is preferred that the inlet curvature 60 variation be sinusoidal. Making the inlet curvature 60 smaller at and/or near the wider lip locations 80, as taught herein, does very little to increase the lubricant shear area of the footprint 38, but significantly impacts the size of Dimension B 2 .
  • the local cross-sectional geometry of the inlet curvature 60 governs Dimension A 2 of the footprint 38, and therefore directly influences the size of Dimension B 2 .
  • the inlet curvature 60 is a radius
  • a larger radius would produce a larger Dimension A 2 and a smaller Dimension B 2 .
  • the large inlet radii on seals constructed in accordance with U.S. Patents 6,315,302 and 6,382,634 do not perform as well as desired.
  • the herein-disclosed understanding of the critical relevance of Dimension A 2 and Dimension B 2 is quite contrary to the conventional wisdom that such prior art seals are based on, and as such represents an inventive step.
  • the theoretical intersection 62 is preferably a zig-zag shape modified by small blending curves 68 at the narrower lip locations 82, and by blending curves 69 at the wider lip locations 80, so that the inlet curvature 60 and the lubricant side flank 58 are un-faceted.
  • the first extent line 64 and the second extent line 66 preferably have zig-zag shapes that are generally similar to that of the theoretical intersection 62.
  • the zig-zag shapes of the first extent line 64 and the second extent line 66 are preferably blended by curves at the wider Hp locations 80 and at the narrower lip locations 82.
  • the second extent line 66 is blended by blending curves 84 at the narrower lip locations 82, and by blending curves 86 at the wider lip locations 80.
  • An example of an appropriate curvature basic dimension for the blending curves 68, blending curves 69, blending curves 84 and blending curves 86 would be a radius in the range of 0.050" to 0.200", and preferably in the range of about 0.100" to 0.150".
  • Another example would be that these blending curves should have a curvature basic dimension no looser than that of a 0.200" radius, and preferably no looser than that of a 0.150" radius.
  • a 0.250" radius would be considered to be a looser curvature than a 0.200" radius, and a 0.100" radius would be a tighter curvature than a 0.200" radius.
  • the term "basic dimension” has the same definition as is given by Section 1.3.9 of ASME Yl 4.5M- 1994 "Dimensioning and Tolerancing.”
  • the lubricant side flank 58 preferably varies in slope about the circumference of rotary seal 10. Referring to FIGS. 2B-2D, the slope of the lubricant side flank 58 is represented by angle ⁇ . As can be seen in FIGS. 2B-2D, the slope of the lubricant side flank 58 is preferably steeper at and/or near the wider lip locations 80, and less steep at and/or near the narrower Hp locations 82.
  • lubricant side flank 58 can be curved or straight, or a combination of straight and curved portions, when viewed in a cross-section aligned with the theoretical axis of rotary seal 10 (such as the illustrations of FIGS. 2B-2D).
  • the lubricant side flank 58 could be a curve that varies in slope.
  • One possibility is to utilize a curve that varies from a given radius at the narrower lip locations 82, to an infinite radius (e.g., a straight line) at and/or near the wider lip locations 80.
  • the difference between a line and a curve is insignificant due to the relatively small size of the lubricant side flank 58.
  • shape, or a combination of the two can be used to achieve the desired result.
  • the core idea is that the slope of the lubricant side flank 58 changes, being steeper near and preferably at the wider lip locations 80, and less steep nearer and preferably at the narrower lip locations 82. This is in contrast to the teaching of U.S. Pat. 6,685,194; for example, see column 15, lines 27-35 of the '194 patent. It is preferred that the variation in the slope of the lubricant side flank 58 be sinusoidal.
  • the first footprint edge L of the footprint 38 may be defined by either the lubricant side flank 58 or by the inlet curvature 60.
  • the least slope of the lubricant side flank 58 and the larger size of the inlet curvature 60 near the narrower lip locations 82 a hydrodynamically efficient, gradual convergence exists between the dynamic lip 16 and the relatively rotatable surface 8 in the region that is circumferentially aligned with the EWH dimension, regardless of whether the first footprint edge L happens to be defined by the lubricant side flank 58 or by the inlet curvature 60.
  • the dynamic lip 16 can be wider (for more sacrificial material to accommodate third body wear) and the size difference between Dimension B 2 and Width Wi can be maximized, while still fitting within a seal overall width that is compatible with the groove designs and any spring designs present in existing equipment. This enables the seal of the present invention to easily retrofit into existing equipment.
  • the slope of the lubricant side flank 58 steeper at and/or near the wider lip locations 80, void volume within the gland is conserved, making the rotary seal 10 more compatible with skew-resisting confinement in severe operating conditions.
  • the varying slope of the lubricant side flank 58 helps minimize the volume of the rotary seal 10 in order to assure sufficient void volume within the gland to accommodate tolerances, seal thermal expansion, seal material displaced by compression, and swelling. This in turn helps to maintain interfacial contact pressure within a range that is compatible with efficient hydrodynamic lubrication, while accommodating a relatively large dynamic surface 56 width at the widest locations of the dynamic lip 16, which ensures that the value of EWH remains positive and effective in severe service conditions.
  • the shallow slope of the lubricant side flank 58 at the narrower lip locations 82 provides those portions with more stiffness and support, compared to the situation that would occur if the steeper slope at the wider lip locations 80 were also used at the narrower lip locations 82.
  • This stiffness allows the narrower lip locations 82 to better resist conditions where the pressure of the second fluid 14 is greater than that of the first fluid 12, such as during down-hole swab events.
  • the additional stiffness has several benefits. It minimizes displacement of the second footprint edge E] relative to Location P 2 , preserving lubrication. It reduces distortion of the dynamic exclusionary intersection 44, facilitating exclusion of the second fluid 14. It also makes the rotary seal 10 moderately less susceptible to circumferential compression-induced buckling, twisting and skewing in applications where the rotary seal 10 is not held in skew-resisting confinement.
  • the footprint 38 has widest locations 70 and narrower locations 72.
  • a zone of increased contact pressure termed herein as a "lubricant-side pressure ridge" exists near the first footprint edge L.
  • the center of the lubricant- side pressure ridge is schematically illustrated by ridgeline 74.
  • This ridgeline 74 is representative of the location of the peak contact pressure at any specific circumferential position along the lubricant-side pressure ridge.
  • the contact pressure varies from a maximum value at ridgeline 74 to zero at first footprint edge L.
  • the contact pressure along the lubricant-side pressure ridge may vary from Location P 2 to Location Pi (a) and from Location P 2 to Location Pi (b > the variable inlet curvature 60 of the seal causes the contact pressure at Location Pj (a) and Location Pj (b) to be much closer in value to that at Location P 2 , and significantly less than the contact pressure in the prior art at Location Pj of FIG. 1.
  • the contact pressure at Location Pj (a) and Location Pj (b) By so-lowering the contact pressure, lubrication of the mating surfaces of the rotary seal 10 and the relatively rotatable surface 8 in the vicinity of Location Pj (a) and Location P 1 (J3) is greatly enhanced.
  • Location P 1 ( ⁇ and Location P] (b) can be thought of as "reversing locations" on ridgeline 74.
  • the ridgeline 74 is preferably a generally zig-zag shape as shown, comprising more or less straight lines blended by small joining curves 73 and small joining curves 75. This shape provides the lubricant-side pressure ridge with the advantage, as compared to the prior art, of not being generally circumferential over a significant circumferential distance at and near the narrower locations 72 of the footprint 38.
  • This improved orientation allows better lubrication of the mating surfaces of the rotary seal 10 and the relatively rotatable surface 8 in the vicinity of Location Pi (a) and Location Pi ( t >) -
  • the keeping of the generally circumferentially oriented portions of the ridgeline 74 as short as practicable via use of the small joining curves 73, without creating a facet on the inlet curvature 60 or on the lubricant side flank 58 while keeping the interfacial contact pressure relatively low, is novel.
  • An example of an appropriate curvature basic dimension for the joining curves 73 and the joining curves 75 would be a radius in the range of 0.050" to 0.200", and preferably in the range of about 0.100" to 0.150".
  • Other curve shapes are possible, but preferably they would have a curvature basic dimension no looser than that of a 0.200" radius, and preferably no looser than that of a 0.150" radius.
  • Varying the size of the inlet curvature 60 to be larger at and/or near the narrower lip locations 82 is quite unconventional in view of the abrupt restrictive diverter teachings of U.S. Pat. 6,109,618.
  • the present invention makes portions of the trailing edges of the waves less abrupt, allowing even more lubricant to escape at the wave trailing edges.
  • the effectiveness of this approach was certainly not obvious prior to testing, and the excellent results were contrary to prior engineering judgment.
  • the brilliance of the variable inlet curvature 60, as taught herein, is that
  • the value of EWH remains positive and effective throughout the useful temperature range of the material used in the construction of rotary seal 10.
  • the EWH dimension is sized such that, unlike the prior art, it remains present (having a positive value rather than a negative value) and functional throughout the useful elevated temperature range of the polymer used in the construction of the rotary seal 10, even if used with skew-resisting constraint. Based on the conventional wisdom, the former ineffective design methodology was to evaluate the footprint wave height in extreme elevated temperature conditions.
  • the current seal design methodology which is an aspect of the present invention, is to evaluate the EWH dimension in extreme elevated temperature conditions that represent the upper limit of the useful temperature range of the polymer, to verify that the value of EWH remains positive throughout the useful temperature range of the material used in the construction of the seal, and preferably to insure that the value of EWH remains greater than or equal to 0.020".
  • This methodology is best implemented via computer simulations utilizing three dimensional large displacement finite element analysis modeling that incorporates a conservative linear coefficient of thermal expansion assumption of 13 X 10-5 inches per inch per degree F, and takes into account the effects of skew-resisting confinement.
  • Downhole drilling equipment is ordinarily cooled by the circulating drilling fluid to a temperature that is lower than the local environment.
  • the temperature of the equipment may reach the temperature of the surrounding geological environment, if allowed to soak long enough without circulation.
  • the useful elevated temperature range used in seal design should be considered to be the temperature that the polymer can withstand for brief periods of time so long as adequate lubrication is present.
  • This "abuse temperature" is greater than the typically quoted long-term temperature capability of a material.
  • hydrogenated nitrile typically is given an extended-term temperature exposure rating of about 300 0 F, but in a seal constructed of such material and employed in accordance with U.S. Patent 6,315,302, it is highly desirable that the value of EWH remain positive if the seal is temporarily exposed to an "abuse temperature” that is 50°F higher than that, and preferably the value of EWH will remain greater than or equal to 0.020".
  • TFE/P The service rating of TFE/P is typically 450 0 F, and the service rating of FKM is typically 400 0 F.
  • FKM the service rating that typically governs the seal design. Therefore, a TFE/P -FKM composite seal would preferably have a service rating of 400 0 F, and in a spring-loaded seal constructed of that material combination, it would be desirable that the value of EWH remain positive if the seal is temporarily exposed to an "abuse temperature" that is 50°F higher than that, and preferably the value of Ey /H will remain greater than or equal to 0.020".
  • the validation temperature used in the computer model of the seal when validating the seal design for a positive EWH value, should be at least equal to the operating temperature limit of the least temperature-capable elastomer used in the construction of the seal (as that operating temperature limit is generally understood within the elastomer industry), and preferably the validation temperature should be 50°F greater than the aforesaid operating temperature limit.
  • Dimension B ⁇ 3) and Dimension Bj (b) are preferably different in size, so that Location P 1 ⁇ ) and Location Pi (b) are misaligned — i.e., offset — by Offset Dimension X.
  • This offset is desirable so that the film disturbances created by the direction reversals of the ridgeline 74 at Location Pi (a) and Location Pj (b) do not lie in one- another's wake. This minimizes the circumferential extent of each such film disturbance and facilitates lubrication.
  • the relative size of Dimension Bi (a) and Dimension Bj (b) is preferably controlled by the local cross-sectional geometry of the inlet curvature 60 of the dynamic lip 16 that is shown in FIGS. 2-2D.
  • the size of the inlet curvature 60 differs between Section 2B-2B and Section 2D-2D; this causes Dimension Bj (a) and Dimension Bj (b) to differ at the corresponding locations of the footprint 38.
  • the salient point is that Location Pi (a) and Location Pi ⁇ ) are offset with respect to each other, regardless of how the offset is achieved.
  • Another way of saying this is that at least some of the joining curves 73 are misaligned with respect to others.
  • some of the waves of the ridgeline 74 are different than other of its waves.
  • the cooperative benefits of the various features provides more complete lubrication, especially in the un-swept zone, in either direction of rotation.
  • the invention is suitable for a wider range of service conditions, including faster and slower rotary speeds, higher differential pressures, and thinner lubricants.
  • Running torque is reduced, resulting in less self-generated heat.
  • the result is better tolerance to high ambient environment temperature, less heat-related compression set, less footprint spread, less seal wear, longer polymer life, a higher retained modulus for improved extrusion resistance, lower interfacial contact pressure when installed in skew-resisting confinement, less slippage within the groove, and less tendency to cause floating compensation pistons to rotate.
  • FIGS. 2-2D includes several desirable features that are most advantageously used together, however simplifications are possible where one or more of the features are omitted or revert to the teachings of the prior art.
  • FIGS. 3, 4, 5, 6, 7 and 8 represent simplifications and alternate embodiments of the invention, and are fragmentary views of a rotary seal 10 in the uncompressed condition thereof, showing a seal that is relatively large or infinite in diameter, or as a smaller seal would appear if a short portion thereof were forced straight, so that no curvature-related foreshortening is apparent.
  • Cutting planes 2B-2B, 2C-2C, 2D-2D, and 4B-4B correspond to the cross-sections shown in FIGS.
  • FIGS. 3, 4, 5, 6, 7 and 8 various previously defined features of the rotary seal 10 are labeled in FIGS. 3, 4, 5, 6, 7 and 8, such as the first seal end 34, second seal end 36, theoretical intersection 62, first extent line 64, second extent line 66, dynamic surface 56, lubricant side flank 58, inlet curvature 60, wider lip locations 80, narrower lip locations 82, blending curves 84, blending curves 86, First Reversing Location Rj, Second Reversing Location R 2 and Offset
  • the rotary seal 10 of FIG. 3 is a simplification of the preferred embodiment of the present invention that differs from that of FIG. 2A in one respect — each of the narrower lip locations 82 (at cutting planes 2D-2D) is substantially the same, which means that Offset Dimension T equals zero.
  • the inlet curvature 60 varies in curvature about the circumference of the seal, being a tighter curve at cutting plane 2C-2C, and looser at cutting plane 2D-2D.
  • the rotary seal 10 of FIGS. 4 and 4B is a simplification of the preferred embodiment of the present invention that differs from that of FIG. 2 A in one respect —the inlet curvature 60 does not vary between cutting plane 2C-2C and cutting plane 4B-4B, as taught by the prior art.
  • First Reversing Location R ⁇ is offset from Second Reversing Location
  • FIG. 4B which is representative of the cutting plane 4B-4B in FIG. 4, various previously defined portions of rotary seal 10 are labeled for orientation purposes, such as angle ⁇ , dynamic lip 16, first seal end 34, second seal end 36, dynamic exclusionary intersection 44, static sealing surface 46, first material layer 48, second material layer 49, projecting static lip 54, lubricant side flank 58, inlet curvature 60 and theoretical intersection 62.
  • the inlet curvature 60 in FIG. 4 is the same as in the view of FIG. 2C, while the slope of the lubricant side flank 58 is shown as being the same as in FIG. 2D.
  • the rotary seal 10 of FIG. 5 is a simplification of the preferred embodiment of the present invention that differs slightly from that of FIG. 2 A in that the theoretical intersection 62 is sinusoidal, as taught by the prior art.
  • the variable size of inlet curvature 60 that is taught in this specification can be used with various wavy shapes, however it is best employed with the modified zig-zag wave shape of FIG. 2 A.
  • the rotary seal 10 of FIG. 6 is a simplification of the preferred embodiment of the present invention that differs slightly from that of FIG. 3 in that the theoretical intersection 62 is sinusoidal, as taught by the prior art.
  • the rotary seal 10 of FIG. 7 is an alternate embodiment of the present invention that differs slightly from that of FIG. 4 in that the theoretical intersection 62 is sinusoidal, as taught by the prior art.
  • the first extent line 64 and the second extent line 66 have shapes that are similar to that of the theoretical intersection 62.
  • the rotary seal 10 of FIG. 8 is an alternate embodiment of the present invention that accomplishes the Offset Dimension T between First Reversing Location Rj and Second
  • Reversing Location R 2 by having a different dynamic lip width at the First Reversing Location Ri, compared to the lip width at the Second Reversing Location R 2 .
  • some of the narrower lip locations 82 have a different width than other of the narrower lip locations 82, and in general, some of the waves of the dynamic lip 16 are different than other of its waves.
  • some of the waves of the second extent line 66 are different than other of its waves.
  • the inlet curvature 60 may, if desired, be the same size throughout, as shown. Note that some of the blending curves 68 of the theoretical intersection 62 are offset with respect to other of the blending curves 68; this means that some of the waves of the theoretical intersection 62 are different than other of its waves.
  • FIGURES 3 A, 4A, 5 A, 6A, 7 A and 8 A represent the footprint 38 of the simplifications and alternate embodiments that are shown in FIGS. 3, 4, 5, 6, 7 and 8, respectively.
  • various previously defined portions of the footprint 38 are labeled in FIGS. 3A, 4A, 5A, 6A, 7A and 8A, such as Dimension Aj (a) , Dimension Ai (b) ,
  • Width Wi The previously defined Width Wi is labeled in FIGS. 3A, 4A, 5A, 6A, and 7A to orient the reader; in FIG. 8A, two different sizes of narrower width location are shown as Width Wi(a) and Width W 1 (TD).
  • the peak pressure of ridgeline 74 at the narrowest points of the footprint 38 are circumferentially aligned, unlike those in FIG. 2E.
  • Dimension B] (a) and Dimension Bj (b) are similarly sized, so that Location Pi (a) and Location Pi (b) are substantially aligned.
  • the footprint 38 of FIG. 4A differs from that of FIG. 2E in one important respect — in FIG. 4 A the contact pressure is relatively high at and near Location Pi (b) because the inlet curvature 60 in FIG. 4 does not vary between cutting plane 2C-2C and cutting plane 4B-4B.
  • the peak pressures of the ridgeline 74 at the narrower locations 72 of the footprint 38 are circumferentially misaligned — i.e., offset — by Offset Dimension X so that the film disturbances created by the direction reversal of the ridgeline 74 at Location P] (a) and Location Pj (b) do not lie in one-another's wake.
  • Some of the waves of the ridgeline 74 are different than other of its waves.
  • the graph of FIG. 7B represents interfacial contact pressure at selected circumferential slices of the rotary seal 10 of FIGS. 2A and 7.
  • the slice representative of the rotary seal 10 of FIG. 2 A was taken between cutting planes 2C-2C and 2B-2B, and the slice representative of the rotary seal 10 of FIG. 7 was taken between cutting planes 2C-2C and 4B- 4B.
  • One circumferential slice is aligned with Location Pi (a) of FIG. 2E, and the other circumferential slice is aligned with Location Pi (b) of FIG. 7 A.
  • the gradient and magnitude of the FIG. 2E contact pressure is preferable to that of FIG. 7A. This shows the benefit of the wave pattern and the varying size of the inlet curvature 60 of the rotary seal 10 of FIG. 2 A, compared to that of FIG. 7.
  • FIGURE 8 A is representative of the footprint 38 of the rotary seal 10 that is shown in FIG. 8.
  • the peak pressures of the ridgeline 74 at the narrower locations 72 of the footprint 38 are circumferentially misaligned — i.e., offset — by Offset Dimension X so that the film disturbances created by the direction reversal of the ridgeline 74 at Location Pi (a) and Location Pj ⁇ ) do not lie in one- another's wake.
  • Offset Dimension X is governed by the Offset Dimension T and the local cross-sectional geometry of the inlet curvature 60 that is illustrated in FIG. 8.
  • the un-swept zone is defined by Width W](a), since it is smaller than Width Wj (b) .
  • Some of the waves of the ridgeline 74 are different than other of its waves, and some of the waves of the First Footprint Edge L are different than other of its waves.
  • FIGURE 9 is a fragmentary cross-sectional view that provides a general overview of another preferred embodiment of the present invention.
  • FIGURE 9A shows the rotary seal 10 of FIG. 9 in its uncompressed condition. Many of the previously described features are numbered in FIGS. 9 and 9 A to orient the reader.
  • Machine 2 incorporates a first machine component 4 and a second machine component 6 that includes a relatively rotatable surface 8.
  • the rotary seal 10 has a generally circular, ring-like configuration and at least one dynamic lip 16 that is also generally circular in form. At least a portion of the dynamic Hp 16 is held in compressed, contacting relation with the relatively rotatable surface 8, and establishes sealing engagement with the relatively rotatable surface 8, to retain a first fluid 12, to partition the first fluid 12 from a second fluid 14, and to exclude the second fluid 14.
  • the relatively rotatable surface 8 has relative rotation with respect to the dynamic lip 16 and with respect to first machine component 4.
  • the dynamic lip 16 incorporates a dynamic exclusionary intersection 44 of abrupt substantially circular form that is substantially aligned with the direction of relative rotation.
  • the rotary seal 10 is oriented by the first machine component 4, which has a generally circular seal groove that includes a first groove wall 18 and a second groove wall 20 that are preferably in generally opposed relation to one another.
  • the first machine component 4 also has a peripheral groove wall 22 that is located in spaced relation to the relatively rotatable surface 8, and compresses the rotary seal 10 against the relatively rotatable surface 8.
  • first groove wall 18 and second groove wall 20 are shown to be in fixed, permanent relation to one another, such is not intended to limit the scope of the invention, for the invention admits to other equally suitable forms.
  • first groove wall 18 and/or second groove wall 20 could be configured to be detachable from the first machine component 4 for ease of maintenance and repair, but then assembled in more or less fixed location for locating and constraining the rotary seal 10.
  • the rotary seal 10 is held in skew-resisting confinement by virtue of simultaneously contacting the first groove wall 18 and the second groove wall 20 under operating conditions, hi the prior art axially-constrained seals made in accordance with U.S. Pat. 6,315,302, even though the designers meticulously evaluated the footprint wave height in elevated temperature conditions, the heretofore unknown E ⁇ VH dimension became compromised in elevated temperature conditions, causing loss of lubrication in the un-swept zone.
  • This and other previously described lubrication problems are managed by incorporating any, or all, of the design features that were previously disclosed in conjunction with FIGS. 2-8.
  • Rotary seal 10 defines a first seal end 34 that generally faces the first groove wall
  • Rotary seal 10 also defines a second seal end 36 that generally faces the second groove wall 20 and the second fluid 14.
  • the compression of the dynamic lip 16 against the relatively rotatable surface 8 establishes and defines an interfacial contact footprint, shown generally at 38, between the dynamic lip 16 and the relatively rotatable surface 8.
  • the footprint 38 has a first footprint edge located generally at L and facing the first fluid 12, and has a second footprint edge located generally at E and facing the second fluid 14.
  • the second footprint edge E is established by compression of the dynamic exclusionary intersection 44 against the relatively rotatable surface 8.
  • the footprint 38 can take the form of any of the footprints shown in FIGS. 2E, 3A, 4A, 5A, 6A, 7A or 8A, with the same characteristics and benefits. The labels in those figures are therefore appropriate to this discussion of FIGS. 9 and 9 A.
  • first seal end 34 may be wavy and vary in position relative to the second seal end 36, as taught by U.S. Pat. Appl. Pub. 2007/0205563.
  • This embodiment, and other embodiments, may if desired also incorporate an exclusion edge chamfer in accordance with the teachings of U.S. Pat. 6,120,036.
  • the dynamic lip 16 and the footprint 38 can have any or all of the attributes previously described in conjunction with FIGS. 2 to 8A; the following citations of such attributes do not represent an exhaustive list of the previously described attributes.
  • the theoretical intersection 62 can have the previously described modified zig-zag shape, or other desired wave shape.
  • the slope of the lubricant side flank 58 can vary around the circumference of the rotary seal 10, being steepest at the widest portions of the dynamic lip 16 to conserve void volume, and to maximize the body length 76.
  • the slope is represented by angle ⁇ , however as described previously, the lubricant side flank 58 can be straight or curved in the cross-sectional view shown (even if straight in the uncompressed condition, the lubricant-side flank 58 tends to become curved in the compressed condition).
  • the varying slope of the lubricant side flank 58 allows the EWH dimension of the footprint 38 to be increased without increasing the volume of the rotary seal 10, compared to the prior art.
  • the varying slope also tends to strengthen the narrowest parts of the dynamic lip 16. This helps the dynamic exclusionary intersection 44 to remain more circular when the pressure of the second fluid 14 is greater than the first fluid 12.
  • the unsupported length (from first groove wall 18 to first footprint edge L) of the rotary seal 10 acts as a skew-resisting spring.
  • the varying slope of the lubricant side flank 58 serves to maximize the exposed length 76 of the body 78 near the wider parts of the dynamic lip 16, which lowers the effective spring rate of the seal body 78 by minimizing the spring force contribution of the dynamic lip 16.
  • FIG. 13 of U.S. Pat. 6,334,619 also discloses a seal with a variably angled flank, the purpose was not to conserve void volume or to preserve body length 76, since that flank is in intimate contact with a similarly shaped backup ring, which supports and constrains the flank, and the flank extends to the end of the seal body.
  • the lubricant side flank 58 is part of an "unconstrained geometry" as taught by U.S. Pat. 6,315,302, so that additional compression or thermal expansion of the rotary seal 10 is compensated by displacement by the unconstrained geometry of the lubricant side flank 58; see col. 11, lines 42- 46 and col. 12, lines 13-20 of the '302 patent.
  • the inlet curvature 60 can vary about the circumference of the rotary seal 10, being a tighter curvature at the wider portions of the dynamic lip 16, and a looser curvature at the narrowest portions of the dynamic lip 16.
  • FIGURE 10 is an enlarged fragmentary shaded perspective view of an uncompressed seal of an embodiment of the present invention in an uncompressed state, as such a seal would be configured for radial compression against a relatively rotatable shaft.
  • This figure is included to facilitate the reader's understanding of the variable slope of the lubricant side flank 58, the variable curvature of the inlet curvature 60, and the generally circular configuration of the dynamic exclusionary intersection 44.
  • the inlet curvature 60 is a radius that is smaller at the widest part of the dynamic lip, and larger at the narrower part of the dynamic lip.
  • the lubricant side flank 58 is a sloped surface that is steeper at the widest part of the dynamic lip, and less steep at the narrower part of the dynamic lip.
  • the wave form is a modified zig-zag shape of the type shown between cutting planes 2C-2C and 2D-2D in FIGS. 2A, 3 and 4.
  • the rotary seal 10 in FIG. 10 is illustrated as being constructed from a single elastomeric material.
  • FIGURE 11 is a fragmentary cross- sectional view of an installed, ring-shaped hydrodynamic rotary seal 10 embodying the principles of the present invention, and installed in the machine that is shown generally at 2.
  • the machine 2 includes a first machine component 4 and a second machine component 6 that defines a relatively rotatable surface 8.
  • FIGURE 11 differs from that of FIG. 2 in that the first groove wall 18 and the second groove wall 20 are both defined by the first machine component 4 and are not both in contact with the rotary seal 10.
  • the arrangement shown in FIG. 11 lacks the backup ring 24, spring 28, and retainer 30 that are provided in FIG. 2.
  • the peripheral groove wall 22 compresses the dynamic lip 16 of the rotary seal 10 against the relatively rotatable surface 8 of the second machine component 6, establishing a footprint shown generally at 38 that has Width W, a first footprint edge generally at L and a second footprint edge generally at E.
  • the static sealing surface 46 has sealed engagement with the peripheral groove wall 22.
  • the second footprint edge E is established by compression of dynamic exclusionary intersection 44 against the relatively rotatable surface 8.
  • the rotary seal 10 is shown in the position it would assume when the pressure of the first fluid 12 is greater than that of the second fluid 14.
  • the second seal end 36 is supported by the second groove wall 20 at all locations except the clearance gap 52.
  • the first seal end 34 is not touching the first groove wall 18.
  • the rotary seal 10 and the footprint 38 can incorporate the features and advantages of the invention that have previously been described, sans the implications of skew-resisting confinement.
  • the rotary seal 10 in FIG. 11 is prevented from skewing and twisting only when the pressure of the first fluid 12 is sufficiently higher than the pressure of the second fluid 14.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Physics & Mathematics (AREA)
  • Fluid Mechanics (AREA)
  • Sealing With Elastic Sealing Lips (AREA)

Abstract

La présente invention se rapporte à un joint d'étanchéité rotatif globalement circulaire qui crée une étanchéité entre des pièces mécaniques relativement rotatives pour la rétention de lubrifiant et l'exclusion d'agents contaminants de l'environnement, et qui incorpore une géométrie de joint d'étanchéité qui interagit avec le lubrifiant pendant une rotation relative afin de distribuer un film lubrifiant dans l'interface d'étanchéité dynamique. Les caractéristiques d'une taille d'entrée variable, d'une inclinaison de flanc de lèvre dynamique variable, et d'une réduction de l'amplitude et de la partie circonférentiellement orientée de la zone de pression à contact interfacial du côté lubrifiant dans la partie la plus étroite de la lèvre, individuellement ou en association, servent à maximiser la lubrification interfaciale dans des conditions de fonctionnement rudes, et servent également à minimiser la zone de cisaillement de lubrifiant, le couple du joint d'étanchéité, le volume du joint d'étanchéité, et l'usure, tout en garantissant une capacité d'adaptation dans les rainures de joint de l'équipement existant.
PCT/US2008/010320 2007-08-31 2008-09-02 Joint d'étanchéité rotatif avec amélioration de la distribution de film WO2009032248A1 (fr)

Priority Applications (1)

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Applications Claiming Priority (6)

Application Number Priority Date Filing Date Title
US96717407P 2007-08-31 2007-08-31
US60/967,174 2007-08-31
US99995707P 2007-10-23 2007-10-23
US60/999,957 2007-10-23
US6741108P 2008-02-28 2008-02-28
US61/067,411 2008-02-28

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Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
ITMI20101296A1 (it) * 2010-07-14 2012-01-15 Bosch Gmbh Robert Gruppo pompa
US10208018B2 (en) 2014-07-02 2019-02-19 Novartis Ag Indane and indoline derivatives and the use thereof as soluble guanylate cyclase activators

Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE322869C (de) * 1920-07-10 Marie Dachsel Geb Schurig Kegelfoermige Dichtungsmanschette
JPS54125276A (en) * 1978-03-23 1979-09-28 Arai Pump Mfg Production of oil seal
US20020175477A1 (en) * 2001-03-07 2002-11-28 Voith Paper Patent Gmbh Sealing device
US6685194B2 (en) * 1999-05-19 2004-02-03 Lannie Dietle Hydrodynamic rotary seal with varying slope

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE322869C (de) * 1920-07-10 Marie Dachsel Geb Schurig Kegelfoermige Dichtungsmanschette
JPS54125276A (en) * 1978-03-23 1979-09-28 Arai Pump Mfg Production of oil seal
US6685194B2 (en) * 1999-05-19 2004-02-03 Lannie Dietle Hydrodynamic rotary seal with varying slope
US20020175477A1 (en) * 2001-03-07 2002-11-28 Voith Paper Patent Gmbh Sealing device

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
ITMI20101296A1 (it) * 2010-07-14 2012-01-15 Bosch Gmbh Robert Gruppo pompa
US10208018B2 (en) 2014-07-02 2019-02-19 Novartis Ag Indane and indoline derivatives and the use thereof as soluble guanylate cyclase activators
US10550102B2 (en) 2014-07-02 2020-02-04 Novartis Ag Indane and indoline derivatives and the use thereof as soluble guanylate cyclase activators

Also Published As

Publication number Publication date
CA2697678C (fr) 2015-06-23
CA2697678A1 (fr) 2009-03-12

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