Abstract:
A double-barrier metal-to-metal seal assembly includes a seal gasket ( 3 ) having an outer flexible seal element ( 6 ) and an inner flat rib ( 8 ). The assembly also includes a pair of hubs ( 4, 5 ) each having an internal machined profile characterised by a tapered profile ( 7 ) which, when the hubs ( 4, 5 ) are mechanically set, provides an interference fit against the outer flexible element ( 6 ) of the seal gasket ( 3 ) so as to provide an outer seal, and an inner profile ( 9 ) comprising a plurality of concentric circular nibs ( 9 ) machined on the hubs ( 4, 5 ) which, when the hubs ( 4, 5 ) are being mechanically set, progressively embed themselves ( 9 ) to a pre-determined depth into the rib ( 8 ) of the seal gasket ( 3 ) to effect an inner seal. The embedment is limited by the face-to-face abutment of the hubs ( 4, 5 ) outside the locus of the outer seal during the mechanical imposition of preload.

Description:
FIELD OF THE INVENTION  
         [0001]    This invention relates in general to pressure containment equipment and in particular to the high-pressure clamp hub and flange connections widely employed in the oil, gas and petrochemical industries.  
         DESCRIPTION OF THE PRIOR ART  
         [0002]    There are several variants of metal to metal seals, the effectiveness of which is dependent upon the degree to which continuous mating of the sealing surfaces is accomplished. The seal is achieved by a combination of:— 
           [0003]    i) making one material relatively soft, normally the ring gasket, such that it deforms under energising load to the surface of the harder material, normally the hub; and  
           [0004]    ii) making the sealing element flex under preload in such a manner that the interference between hub and seal element produces a sealing contact face-to-face further enhanced by the action of the pressure of the fluid being retained.  
           [0005]    Metal-to-metal sealing ring gaskets are mainly type (i) seals whilst type (ii) seals are typified by the renowned Grayloc® seal and certain ring gaskets. For all seals, it is essential that the contact surfaces be in good condition.  
           [0006]    It would be advantageous to have two fully testable metal-to-metal seals in place in a connection for retention of pressurised fluids. In order to be a true double barrier, each seal must demonstrate a full sealing function against pressure originating from the bore. However the outermost seal must be pressure-tested from within and this means that the innermost seal shall be pressurised from behind. The innermost seal must therefore be a fully bi-directional seal or have pressure support in the bore i.e. bore pressurised to balance the test pressure between the two seals. Pressure-balancing the innermost seal by pressurising the bore can be expensive and time-consuming. It can also be dangerous as it would normally require large volumes of fluid under pressure. Furthermore, it would be advantageous to be able to ascertain whether there is pressurised fluid still being retained within the bore prior to disconnection and to ensure that when such pressure manifests itself, it could still be retained even after initial separation of the hubs/flanges of the connection. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0007]    [0007]FIG. 1 depicts a gasket and flange with typical annotation;  
         [0008]    [0008]FIG. 2 depicts a seal and hubs in accordance with the present invention prior to mechanical make-up;  
         [0009]    [0009]FIG. 3 shows a partial section of the hub engaging the seal of FIG. 2 to a magnified scale;  
         [0010]    [0010]FIG. 4 depicts the hubs of FIG. 2 made-up face-to-face with a seal fully engaged; and  
         [0011]    [0011]FIG. 5 depicts a variant of seal gasket in accordance with the present invention with an elastomeric or non-metallic element at the inner seal faces. 
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS  
       [0012]    The type of seal to be employed in any pressure-containing connection is determined after due consideration of pressure, thermal limits, legislative requirements and customer practice and preference. Metal-to-metal seals function by a combination of preload and pressure-assistance. After the hubs within which the seal is located are mechanically made-up by bolting or clamping to full required preload, the seal would be pressure-tested to verify both the mechanical integrity of the connector and the effectiveness of the seal.  
         [0013]    For gasket types such as ring gaskets approved by the American Petroleum Institute (API), any reduction in seal-hub contact stress below an empirical multiple of the pressure of the retained fluid would result in unacceptable leakage. Gaskets have been incorporated into national design standards such as ASME. III and VIII (U.S.A.) and BS 5500 (UK)  
         [0014]    Graph 1 shows some characteristics typical of gaskets under load. As pre-load is applied, the gasket follows broken line O-A which is non-recoverable loading. During this phase, the gasket is forced to conform to the profile of the sealing face of the hub, filling any irregularities and surface asperities. Point A is the minimum load on the gasket that will provide an acceptable seal and is known as the gasket seating stress designated, ‘y’ in applicable BS and ASME Codes. 
         
 
         [0015]    Graph 1—Make-Up Characteristics of Metal Gaskets  
         [0016]    Section A-C is the useful sealing range of the gasket with C being the point at which some structural breakdown of the sealing element occurs. When the connection is made up on assembly stress, B will be imposed on the gasket. As internal pressure builds up, the gasket stress will reduce. Should this load fall below an empirical multiple of the retained fluid pressure then leakage occurs. This multiple, designated ‘m’ in industry Codes, is dependent upon the type of gasket and may be found in tabulated reference ASME VIII Section 1, Boiler and Pressure Vessel Code.  
         [0017]    As all metal—metal seals leak to some extent, the values of ‘m’ and ‘y’ are those known from experience to give acceptable leakage rates.  
         [0018]    Two criteria have been established for gasket design in BS 5500:  
         [0019]    1. Preload must be sufficient to ensure that the gasket seating load is attained; and  
         [0020]    2. Preload must be sufficient to hold internal pressure end load, plus minimum seal contact load.  
         [0021]    Referring to FIG. 1—Flange and Seal Diagram which shows a gasket  1  on the point of being made-up to a flange  2  (mating flange not shown for brevity), the equations for determining minimum preload to meet provision  1  is:  
           W   1   =πbGy   (1)  
         [0022]    where b is the effective gasket width and G is the effective gasket sealing diameter.  
         [0023]    Provision (2) is to ensure that once test pressure P test  is applied to the joint, there is enough preload remaining to hold the sealing interface to a contact stress above the threshold value to effect a seal. The pertinent equation is:  
           W   2   =πG   2   P   test /4+2π GbmP   test   (2)  
         [0024]    The first part of the above equation is the pressure end load which, of course, must be resisted by the pre-load. The second part is the minimum sealing load. There is a factor of 2 included over and above factor ‘m’. This is because effective gasket sealing width is an empirical concept, based upon the supposition that the hub pivots about the outer edge of the gasket and it is this outer area which forms the seal width.  
         [0025]    Preload applied to any joint will be at least the greater of equations (1) and (2). Modern tools such as Finite Element Analysis (FEA) should be employed where possible to assess the effects on seal loading due to bending moments.  
         [0026]    Refer now to FIG. 2 which shows the seal or ring gasket  3  of the invention landed in a profiled hub  4  with the mating hub  5  ready to be connected. The gasket  3  has an outer element  6 , similar to that of the classic pressure-energised seal as exemplified by the Grayloc®, which shall provide an outer metal-to-metal seal of the pressure-energised type as it is energised by, and set to seal against, a smooth, tapered surface  7  on each hub. The rib of the gasket  8  shall be used to form the inner seal.  
         [0027]    [0027]FIG. 3 shows the hubs  4 , 5  fully mated and a resultant double metal-to-metal seal system. The hubs  4 , 5  have a series of concentric toothed profiles termed nibs  9  directly opposite the rib  8  of the gasket  3 . The hubs  4 , 5  have a material yield strength preferably approximately 2.5 times greater than that of the gasket  3  in order that the nibs  9  shall embed themselves into the rib  8  of the gasket  3 .  
         [0028]    Referring to Graph 2 which shows the normal stress versus embedment depth of a steel wedge of 520 N/mm 2  (75 000 psi) yield strength being driven into a steel plate of 210 N/mm 2  (30 000 psi) yield strength, it is seen that an embedment depth of 0.75 mm (0.030 in) can be attained at which point the contact stress reaches the yield strength of the wedge. Any attempt to drive the wedge further into the plate would cause the wedge to yield and plastically deform. 
         
 
         [0029]    It is preferable that the nibs  9  do not undergo the permanent deformation which would occur were the normal contact stress between the rib  8  of the gasket  3  and each of the hubs  4 , 5  to exceed the yield stress of the hub material on which the nibs are machined. This is to ensure full repeatability of embedment geometry and make-up conditions. However, if the nibs  9  were to be subject to contact stresses in excess of their material yield point, the nib material would work-harden locally which would not be detrimental to the repeatability of sealing action provided that the geometry of the nibs  9  does not change beyond preferred limits.  
         [0030]    It is further preferable that the outer element  6  of the gasket  3  make contact with its mating surface  7  on the hubs  4 , 5  before the rib  8  touches the nibs, both as the gasket  3  is placed in one hub  4 , and as the other hub  5  is mated with the first hub  4  and the gasket  3 . This ensures that the gasket  3  shall be centralised in the assembly of the connection. It is anticipated that the tapered surface  7  within the hub profile on which the outer element  6  of the gasket  3  engages and is compressed as hubs  4 , 5  are subjected to progressive mechanical preloading need not necessarily be the ultimate sealing face of the outer seal; the sealing face could be a parallel bore or a reduced taper sited immediately behind the energising tapered surface  7 .  
         [0031]    The nibs  9  have a section which is preferably symmetrical with a flat contact face of a width preferably circa 0.1 mm to 0.2 mm and preferably of an included angle of 60-80°. The quantity of nibs is dependent upon the application but should preferably be 3-4 because, whilst it is preferable that the local contact stresses not exceed the yield strength of the nib/hub material, a compressive stress in the axial direction should be produced in the drive rib  10  preferably of 30-60% of the hub/nib yield strength. This compressive axial stress further imparts compressive radial and hoop stresses to the seal rib  8  and the drive rib  10  through Poisson&#39;s effect. The use of one nib only combined with an inner elastomeric or non-metallic seal is an option.  
         [0032]    The hubs  4 , 5  are preferably symmetrical and sized such that when the nibs  9  are achieving the preferred embedment into the seal rib  8  whereby the resulting contact stress of nib to seal rib is approaching the value of the nib/hub yield stress, the hubs  4 , 5  make face-to-face contact and all further preload goes through the hub-to-hub contact face  11 . This ensures that there is no excessive contact stress between nibs  9  and gasket rib  8  and that bending loads are accommodated on the hubs  4 , 5 . The embedment distance for a given seal material can be determined from its own specific, empirical graph of contact stress versus embedment depth.  
         [0033]    [0033]FIG. 4 shows the gasket  3  between the mating hubs  4 , 5 . The flexible element  6  is first to be energised. As the hubs  4 , 5  are progressively tightened, nibs  9  dig into the relatively soft rib  8  effecting a series of seals close to the bore. Nibs  9  are a series of concentric planar teeth whose yield strength is preferably 2-2.5 times that of the ring gasket material.  
         [0034]    It is evident that the seal thus created can be tested in the void  12  between the nibs and the outer element through ports  13  linking the external source of test pressure to said void  12 . Pressure-balancing holes  14  through a non-sealing diameter on the gasket rib  8  ensure completeness of the test.  
         [0035]    An essential feature is that the outer profile  6  should centralise the gasket  3  in a first hub  4  prior to make-up to a second hub  5 . Were this not so, the gasket would be eccentric and lock in this position when the nibs bite and grip. The outer profile may then not achieve the required seal. Said centralisation is achieved by ensuring that the length of the outer seal element is such that, when the gasket  3  is placed in a hub  4 , the outer seal element  6  contacts its mating tapered sealing face  7  on the hub  4  before the nibs  9  contact the rib  8  indeed, it is only during the mechanical make-up that the nibs  9  contact, and embed into, the rib  8  of the gasket  3 .  
         [0036]    The mechanics of forced indentation of a relatively soft material by a harder material in a geometrical form that enhances the indentation requires some explanation even to those experienced in the art of mechanical sealing as it is the founding theory behind the design.  
         [0037]    It has been demonstrated that where shear stresses exist on the sealing interface, the most effective seal occurs where the shear stress is the greatest. This does not necessarily coincide with the area of highest normal stress. When nibs are driven into a flat interface by a normal force, stresses will develop as seals are created. There is an “apparent” seal stress which is determined by dividing the seal load by the total area of the seal interface which, in this case, would be the circumferential area between and including the inner and outer nibs. The actual seal stress is the seal load divided by the actual contact area and is readily calculable as the depth to which the nibs are driven will be known from the hub and seal geometries, i.e. the nibs can only be driven up to the point where the hubs attain face-to-face abutment.  
         [0038]    Prior studies into the effects of driving a relatively hard metal into work-hardening softer metal revealed the following five regimes of metal to metal sealing:  
         [0039]    Regime 1  
         [0040]    As the surfaces make contact under light but increasing load, the peaks of the surface profile of the softer material will yield for, although the contact area is small, the actual seal stress is high. Apparent seal stress is half the yield stress of the softer material. Leakage occurs in this regime.  
         [0041]    Regime 2  
         [0042]    Normal stress increases bringing more contact. The apparent stress is 2-3 times the yield of the softer material, but the actual contact stress may only be the yield stress of the softer material. This sealing regime leaks but is slowing as the contact force increases.  
         [0043]    Regime 3  
         [0044]    The actual stress now exceeds that of yield of the softer material which is on the point of bulk plastic flow.  
         [0045]    The amount of work required to produce greater mating rises enormously and is a function of the strain hardening index of the softer material. The actual contact stress is 2-3 times the yield of the softer material as is the actual contact stress due to the far greater contact area. For this regime, a leakage rate of less than 10 −6  cc/sec of helium at STP isolated at 100 atmospheres has been reported.  
         [0046]    Regime 4  
         [0047]    Further increase in normal stress is causing the softer material to flow in bulk with resultant changes in physical dimensions. Strain hardening of the softer material is increasing its relative hardness to approximate that of the harder material which may itself now be on the point of yielding. Contact stress is 2.5-5 times that of the softer material&#39;s yield. There is no leakage.  
         [0048]    Regime 5  
         [0049]    The softer material has now strain-hardened to the point where increases in stress cause the harder material to yield. Regimes 4 and 5 result in only marginally better mating between seal surfaces for a proportionally far greater increase in stress.  
         [0050]    Clearly regimes 4 and 5 must be avoided as any plastic deformation of the harder surface i.e. the nibs, would have a deleterious effect on the reusability of the hubs. It is the preference to have a minimum yield strength ratio of 2.3:1 for hard and soft surfaces. Where a hub material is of 75 000 psi (517 N mm −2 ) yield strength, then the material of the ring gasket would be around 30 000 psi (207 N mm −2 )  
         [0051]    Control over the development of the normal stresses is derived from an experiment during which a hard wedge of the same profile as the nib was driven into a steel plate of 30 000 psi (207 N mm −2 ) yield strength. The resultant plot of normal contact stress against bite depth is shown in Graph 1. As it is intended not to exceed a normal contact stress of 70,000 psi (483 N mm −2 ) for this particular embodiment, then an indentation of 0.03 inches (0.76 mm) per side shall be sized on the assembly of seal and hubs. It has been determined through testing that, when the nibs  9  are spaced within approximately 0.15 inches (4 mm) of each other, a “proximity effect” is produced whereby work-hardening effects are apparently increased and the depth of embedment is less than that for singular embedment for a given contact stress level. In effect, the slope of graph 2 is steeper for proximity-affected embedment than for singular embedment, and provides a correction for indentation of nib into gasket.  
         [0052]    The development of stresses from the preload during hub make-up is explained hereafter. Consider again FIG. 2 showing the hubs  4 , 5  being brought together. The outer element  6  is partially engaged and the nibs  9  are just touching the rib  8 . There is still a gap of 1.5 mm (0.06 in.) to be closed i.e. 0.75 mm (0.03 in.) per face.  
         [0053]    [0053]FIG. 4 depicts the faces of the hubs  4 , 5  as now being in contact with a normal stress at the interface of the rib  8  and nib  9  of 70,000 psi (483 Nmm −2 ).  
         [0054]    The preload that now exists through the nibs for 5⅛ inch API flanges is:  
           F   pre =70 000  psi×π×D× 4 t   fin    
         [0055]    where t fin =0.05 in (1.25 mm) i.e. the width of the embedment per nib in plan view.  
           F   pre =251,227  lb−f (1117  kN )  
         [0056]    The averaged stress through the section of the drive rib  10  is  
         251   ,   227                   lbf   /     π   4            (       6.1   2     -     5.13   2       )       =     29   ,   379                 psi               1117                                  kN   /     π   4            (       155   2     -     130.3   2       )       =     201.9                 N                   mm     -   2                               
 
         [0057]    The stress in that part of the seal profile that carries the nib loading is higher than elsewhere. This stress is mainly plastic at the areas of nib  9  embedment into the rib  8  but averaged elastic stress is established over the section. The latter is recoverable and holds the seal faces i.e. nibs and seal rib in contact at a pressure in excess of the retained fluid pressure.  
         [0058]    ASME VIII recommends an ‘m’ value of 3 for many applications. It should be borne in mind that the ASME Boiler and Pressure Vessel Codes are intended for containment of pressures lower than that for oilfield equipment. A gasket would be destroyed if such loading as ASME indicates were applied and so, for high pressure containment, reduced values of ‘m’ have been applied successfully.  
         [0059]    Preload in a connection in which this invention is employed may be obtained from a Grayloc®-type clamp, or from bolted flanges utilising threaded elements such as bolts, studs and nuts set to those torque values required by design and conditions of operation. Both methods are well-known to engineers expert in the art. The preload calculations hereafter consider the criteria outlined previously where seal seating loads, pressure end loads and loading factors are examined. Loads for the design case where test pressure is 10,000 psi (68.7 Nmm −2 ) are:  
           W   2   =πG   2   P   test /4+2 πG   1   bmP   test +40,000  lbf (177,959  N )  
         [0060]    Pressure end load+nib seating load+flexible element preload  
         [0061]    where,  
         [0062]    G=diameter of first bore seal=5.316 in (135 mm)  
         [0063]    G 1 =diametral mid-point of nibs=5.6 in (142.2 mm)  
         [0064]    b=0.05 in/nib×4 nibs=0.2 in (5.1 mm) i.e.  4×t   fin    
         [0065]    m=2  
         [0066]    then,  
                 W   2     (   lbf   )     =       (       π   4     ×     5.316   2     ×   10   ,   000     )     +     (     π   ×   5.6   ×   0.2   ×   2   ×   10   ,   000     )     +     40   ,   000                       W   2          (   N   )       =       (       π   4     ×     135   2     ×   68.7     )     +     (     π   ×   142.2   ×   5.1   ×   2   ×   68.7     )     +     177   ,   959                                   
 
         [0067]    giving,  
         [0068]    W 2  (lbf)=221,839+70,336+40,000  
         [0069]    W 2  (N)=986,957+309,478+176,000  
         [0070]    W 2 =332,175 lbf  
         [0071]    W 2 =1,461,570 N  
         [0072]    This value of W 2  is the calculated minimum value of preload required to maintain a seal at internal pressure of 10,000 psi (68.7 Nmm −2 ) for a contact multiplier ‘m’ of 2.  
         [0073]    Another condition that must be met from consideration of safety and good engineering practice is that preload should be greater than the combined pressure end load and seal engagement load that would act upon the flexible element of the seal in the event that the inner sealing arrangement fails. Taking a sealing diameter of 6.2 inches (157.5 mm) and terming the resultant load as W 3  yields:  
         [0074]    W 3  (lbf)=(π/4×6.2 2 ×10,000)+40,000  
         [0075]    W 3  (N)=(π/4×157.7 2 ×68.7)+177,959  
         [0076]    Pressure end load flexible element load  
         [0077]    giving, W 3 =341,907 lbf  
         [0078]    W 3 =1,504,391 N  
         [0079]    It is further foreseen as depicted on FIG. 5 that the immediate junction at the bore to be sealed between the innermost nib and seal interface is a suitable site for a non-metallic element  15  attached preferably by adhesive preferably to the rib  8 . There are two junctions per gasket hence two non-metallic elements per gasket. Each non-metallic element should be of an axial thickness slightly greater than the nominal gap which would exist when the hubs  4 , 5  are fully preloaded. i.e. when the nibs  9  are embedded to pre-designed depth in the rib  8  of the gasket  3 . After complete make-up of the hubs  4 ,  5 , each non-metallic element  15  is compressed and may provide seal against the innermost nib embedded in the rib of the metal gasket. There is no nominal gap between the sealing elements behind the non-metallic element  15  and the latter cannot therefore be extruded through any gap as a result of pressure and fail in the manner in which O-rings typically fail. It should be noted that there exist several proprietary or generic compounds available from suppliers of non-metallic seals and gaskets which would suit the purposes described herein.  
         [0080]    It is considered that the problem of crevice corrosion often encountered in the oil and gas industry may be eliminated by use of the previously described non-metallic element as the crevice which would exist between two hubs and a gasket is instead occupied by a pre-compressed non-metallic element bounded by the rib of the seal and the nib and face of the hub.  
         [0081]    There have been several instances where the bore of pressure-retaining equipment has not been depressurised prior to attempted disconnection of a connector assembly with resultant loss of life and property. As the elements which generate the mechanical preload within the present invention in its clamp or flange configuration are unfastened, the innermost seal relaxes and would permit passage of retained fluids from the bore to the void  12 . The outermost seal does not cease functioning until the hubs  4 , 5  are separated by greater than the combined embedment depth of the nibs  9  into the rib  8  of the gasket  3  thus ensuring that any unreleased pressure within the bore can be retained and shall be detected by pressure-monitoring equipment.  
         [0082]    The present invention has advantages over current art. It is seen that the sealing group generated by embedment of the nibs of the two respective hubs into the rib of the seal acts in a bi-directional manner in direct contrast to the outer flexible element of this gasket and of other art which functions in a uni-directional mode. It is therefore evident that the small void between the outer flexible seal element and the inner compressive embedment group may be subject to a pressure test of full rated pressure of the connection immediately after the respective hubs are mechanically set against each other to the requisite preload. This allows the operator to verify the achievement of a dual metal-to-metal seal at each instance of implementation of a mechanical connection without having to prime the entire pressure-containment system or to take measures to isolate internally those elements requiring test. Given that the pressure-containment systems envisaged as suited to use of this invention include systems of pipework, valves, manifolds, Christmas trees and wellheads, the verification of the integrity of the system&#39;s pressurised connections prior to full test of the fluid retaining volume of the pressurised system provides a cost-saving step as system blowdown (rapid depressurisation) on seal failure is extremely unlikely, and there is an additional metal-to-metal seal system in place. Furthermore, the void between the metal seals is suited to the siting of pressure-monitoring fittings for unexposed observation of the connection during service.  
         [0083]    Owing to the highly corrosive nature of retained fluids, a necessary practice within the oil and gas industry is the overlaying, generally applied by one or more welding processes, of essential sealing areas with Corrosion Resistant Alloys, said alloys being proprietary grades which obtain their desired properties typically from nickel, chromium and molybdenum. Such alloys would be suitable for application to this invention and would preferably be present on the pre-machined hub profiles on both the profile against which the outer flexible element of the gasket seals, and where the nibs shall be produced under the final machining operation. In their as-welded condition, these corrosion-resistant alloys are not necessarily of the requisite hardness prior to the first cycle of embedment in the gasket material. These alloys do, in general, have rapid work-hardening characteristics and would attain the requisite hardness after a deliberate, measured pre-embedment into a softer, sacrificial ring material, said pre-embedment intended to bring the nibs up to the requisite hardness prior to operational use with a seal gasket. The sacrificial ring would be typically of soft iron and the nib profile would preferably be modified to take account of the work-hardening pre-embedment operation.  
         [0084]    Such a pre-embedment operation could also be employed on hubs of steels of yield strength less than the preferred ratio of 2-2.5 times that of the seal gasket in order that the local yield strength of the hub material at the nibs be increased by work-hardening to attain the preferred ratio.  
         [0085]    Implementation of a variant of this invention which would allow mechanical make-up in a subaqueous environment encounters the central problem that the water which would be present around the sealing elements is effectively incompressible. In the isolated cavities between the nibs, such as would normally be created by embedment into the rib of the gasket, would be generated great local pressures acting against the mechanical setting force thereby rendering establishment of a secure seal particularly difficult. This problem may be countered by the placement during hub manufacture of a more readily compressible elastomeric material into the troughs of the nibs to create a generally flat profile across the area of the nibs. It is evident that, as the nibs would embed into the rib of the seal, the greatly reduced presence of water within the cavities between the nibs and the presence of a more compressible fluid—the elastomeric material—would facilitate the achievement of a functioning seal. The water trapped in the void between the primary seal elements would be vented through one of the test ports.  
         [0086]    Modifications and improvements can be made to the embodiments hereinbefore described without departing from the scope of the invention.