Threaded joint for steel pipes

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DESCRIPTION

Threaded joint for steel pipes.

Technical field

The present invention relates, in general, to a threaded joint for use in the connection of steel pipes, such as oil well pipes or OCTG (oil exploration and exploitation tubular goods) including pipes and housings, riser pipes and pipelines. for use in the exploration and production of oil wells and gas wells and, particularly, to a threaded joint for steel pipes that has excellent stability with respect to external pressure and compressive strength.

Previous technique

Threaded joints are widely used to connect steel pipes, such as OCTG and riser pipes, for use in industrial oil production equipment.

In the past, conventional threaded joints specified by API (American Petroleum Institute) standards were typically used to connect steel pipes, for use in oil and gas exploration and production. However, in recent years, the environments in which the exploration and production of crude oil and natural gas take place have become increasingly severe, so that special high-performance threaded joints, called premium joints, have become more severe. are using more and more.

A premium gasket typically comprises, on each pipe, a tapered thread, a metal-to-metal seal portion, that is, a sealing surface that has the ability to form a seal when in intimate contact with the metal-to-metal seal. of the other member of the joint, and a protruding torque part, that is, a protruding surface that functions as a stop element during the constitution of the joint.

In the past, since vertical wells were the most common, an OCTG threaded joint could function properly as long as it could withstand a tensile load due to the weight of the pipes attached to it and could prevent leakage of a fluid from high pressure that passes through its interior. In recent years, however, because wells are getting deeper, because the number of directional wells and horizontal wells that have a curved wellbore is increasing, and because of well development in harsh environments, such as offshore or in higher latitudes, has increased, a wider variety of properties are required for threaded joints, such as compressive strength, torsional strength, tightness against external pressure and ease of handling in the field.

Figures 2 (a) and 2 (b) are schematic explanatory views of a premium joint of the ordinary coupling type for OCTG, comprising an externally threaded member 1 (hereinafter referred to as a tip member or simply a tip, and an internally threaded member corresponding 2 (later referred to as cash member or simply cash).

The tip member 1 has, on its outer surface, a male thread 11 and an unthreaded portion 12, called a flange, which is located at the end of the point 1 and is adjacent to the male thread 11. The flange 12 has a metal-to-metal seal part 13 on the outer peripheral surface of the flange and a protruding torque part 14 on the end face of the flange.

The corresponding box member 2 has, on its internal surface, a female thread 21, a metal-to-metal seal part 23, and a torque protruding part 24, which are portions capable of adapting or contacting the thread. male 11, the metal-to-metal seal part 13 and the torque protrusion part 14, respectively, of the tip 1.

Figure 3 is a schematic diagram illustrating the shape and dimensions of a trapezoidal thread, exemplified by an API reinforcing thread. Most of the threads for use in premium gaskets are trapezoidal threads modeled on this API reinforcing thread. Most threads almost directly copy API reinforcing thread dimensions with respect to thread tooth aspect ratio, flank angle, and other characteristics.

In Figure 3, if the thread is an API reinforcing thread that has a thread pitch of 5 TPI (5 threads per inch), for example, the thread height 74 is 1.575mm, the flank angle load 71 is 3 degrees, the angle of the piercing flank 72 is 10 degrees and the axial gap 73 between the piercing flanks is about 100 pm (ie, 30 to 180 pm) on average.

An overlap in the radial direction called interference is provided between the sealing surfaces of the tip and the housing. When the gasket is squeezed until the protruding surfaces of the tip and the box abut each other, the sealing surfaces of the two members are brought into close contact with each other over the entire circumference of the gasket to form a seal.

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The protruding surfaces function as stop elements during tightening, and can withstand almost the entire compressive load applied to the joint. Therefore, they cannot withstand a large compressive load unless the wall thickness of the protruding surfaces is large (or unless the stiffness of the protrusions is high).

When external pressure is applied to a conventional premium gasket, as described above, the applied external pressure penetrates through the gaps between the threads to a portion 31, shown in Figure 2, just before the seal parts.

The wall thickness of a flange is much thinner than that of a pipe body, so it can experience a decrease in radius due to penetrating external pressure. As the external pressure increases, a gap forms between the sealing surfaces, resulting in a leak, that is, a situation where external fluid enters the interior of a pipe.

If a compressive load is applied to a premium joint in situations such as when an OCTG is arranged in a horizontal well or directional well, since most joints have a relatively large gap between the drill flanks, such as the In the case of the API reinforcing thread described above, the threads have a poor ability to withstand compressive loads, so most of a compressive load is borne by the projections.

However, the wall thickness (the load bearing area for a compressive load) of a protruding surface is typically much less than that of a pipe body. Therefore, if a compressive load equivalent to 40 to 60% of the elastic limit of the pipe body is applied, most of the premium gaskets experience considerable plastic deformation of the protruding part of the torque of the box, resulting in a significant reduction in the tightness of adjacent sealing surfaces.

The tightness of a joint with respect to external pressure can be improved by increasing the stiffness of the tip to increase its resistance to radial shrinkage deformation. To this end, a method is often used in which a working process for reducing the diameter of the end of the pipe, called drawing, is previously performed to increase the thickness of the flange wall.

However, if the amount of drawing is too large, in the case of the casing, a pipe that is inserted into the casing may fall into the drawn part, and in the case of a pipe, the drawn part may cause turbulence in a fluid, such as crude oil that flows through the inside of the pipe and cause erosion. Therefore, the wall thickness of the tip flange wall cannot be greatly increased by drawing.

Other conventional techniques for increasing the stiffness of the end of a tip to improve its sealing are described in US Patent No. 4,624,488 and US Patent No. 4,795,200. These patents describe techniques in which the tightness is increased by providing a cylindrical part that does not contact a box at the end of the sealing surface of a tip, such as to increase the rigidity with respect to radial shrinkage deformation of the periphery of the sealing surface of the tip and contact the sealing surfaces of a gasket evenly.

With a pipe joint, even if drawing is done, it is necessary to provide a tapered joint as a sealing surface and a protruding surface within a limited wall thickness. However, in the prior art described above, the protruding surface must be arranged at a location other than the rim because the end of the tip does not bear against the box, whereby the wall thickness of the rim is necessarily reduced.

In this way, there is a limit to the extent to which the stiffness of the flange can be increased, such as to resist a decrease in radius caused by external pressure, and the tightness with respect to external pressure cannot be significantly improved. Furthermore, because the projecting surface cannot be given a sufficient radial width, a high level of compressive strength cannot be achieved and the tightness is poor under a combination of compression and external pressure.

Techniques for giving a thread the ability to withstand a compressive load to improve compressive strength are described in US Patent No. 5,829,797 and US Patent No. 5,419,595, for example. US Patent No. 5,829,797 describes threads in which the bearing flanks and the piercing flanks of trapezoidal threads come into contact with each other and radial gaps are provided both in the troughs of the thread and in the crests of the thread. thread. This thread has a very high capacity to withstand a compressive load because the drilling flanks are always in contact.

US Patent No. 5,419,595 describes a thread in which the gap between the piercing flanks of the trapezoidal threads is reduced to 30 μm or less, so that the piercing flanks contact each other only when a compressive load is applied. Although the ability of this thread to withstand a compressive load is less than that of the thread described in US Patent No. 5,829,797, it is much greater than that of ordinary reinforcing thread.

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However, with the thread described in US Patent No. 5,829,797, if the width of the thread teeth varies, large variations can occur in compressive strength, anti-wear properties, torque constitution, and other properties. Therefore, it is necessary to make the manufacturing tolerances extremely small, and as a result, this thread has a problem that it is not suitable for mass production and is extremely expensive to manufacture.

US Patent No. 5,419,595 has a similar problem. In particular, the gap between the drilling flanks should be set to a value of 0 to 30 m. In this case, the allowable variations in the width of the male thread tooth and the female thread tooth are each only ± 7.5 jum, so that thread cuts become extremely expensive and unsuitable for production in dough.

GB 1,220,856 and JP 58-157087 describe a threaded joint according to the pre-characterization section of claim 1.

US 4,548,431 discloses a threaded sleeve for onshore auger drill pipe members, which has features that increase the amount of torque required to make the connection. The threaded sleeve has both external and internal constitution protrusions. The tip and box are dimensioned so that the face of the box is in contact with the outer shoulder, while there is still clearance between the face of the tip and the inner shoulder. At normal build torque, the outer boss will be tightened to its normal amount. The inner boss will be tightened to its limit only if additional torque is encountered while drilling. The internal protrusion adds to the amount of torque required to produce the threaded sleeve connection.

JP 08-247351 and CN 2366656 Y describe a type of joint coupling for oil well pipes, which includes non-threaded parts to form a seal.

Description of the invention

The object of the present invention is to solve the problems of the prior art described above and to provide a threaded joint for steel pipes which has excellent compressive strength and which greatly improves the overall tightness with respect to external pressure, not only when subjected to external pressure but also when subjected to a combined load of compression and external pressure or tension and external pressure.

The present invention provides a threaded joint for steel pipes comprising a tip and a housing, the tip having a male thread, a sealing surface, and a protruding surface, the housing having a female thread, a sealing surface, and a protruding surface. , the male thread integrating with the female thread, radially interfering the sealing surface of the tip with the corresponding sealing surface of the box and the protruding surface of the tip abutting axially on the corresponding protruding surface of the box, wherein (i) the protruding surface of the tip is provided on a terminal face of the tip; (ii) the tip sealing surface is located on one side of the end of the pipe, close to the male thread; (iii) a front part (15) is provided on the tip, between the sealing surface and the projecting surface, said front part not coming into contact with the part of the box facing the front of the tip, the tapered being of the sealing surface between 5 and 25 degrees and the sealing surface (13) of the tip arranged radially inwards, as measured from the axis of the tip of (i) the tangent (61) to the crest of the male thread and the apex of the tip and (ii) the extension line (62) of the valley of the male thread, characterized because there is a surface discontinuity between the sealing surface and the outer surface of said front part.

Figure 4 is a schematic illustration of a flange used in basic studies conducted in connection with the present invention.

To achieve the object described above, the present inventors performed a finite element analysis with respect to the following four design factors, to determine the shape of a flange, of a premium gasket having the structure described above: (1) the thickness of flange 41, (2) length of flange 42, (3) angle of shoulder 43 and (4) taper of seal (taper of sealing surface) 44. For each factor, they investigated its effect on tightness with respect to external pressure.

As a result, it was found that the thickness of the flange 41 and the length of the flange 42 had the greatest effect on the tightness with respect to external pressure and that the thicker the thickness of the flange and the longer the length of the flange, the more marked it is the improvement of hermeticity with respect to the external pressure.

It was found that to improve the compressive strength, the thickness of the flange is preferably increased as much as possible. As for the angle of the protrusion 43, the tightness under a combined load of compression and external pressure is improved if the angle of the protrusion of a point is such that an external part of the protrusion of a point tapers to form a protruding surface shaped like hook as shown (referred to below as hook angle). However, such a protrusion angle results in greater damage to the protruding surface of a box member through a compression load. Therefore, if a protrusion angle is provided, the protruding surface of the box member is preferably reinforced, or a thread having a high compressive load bearing capacity is preferably used.

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The taper of seal 44 was found to have only a small effect on tightness relative to external pressure. However, if the seal taper is too large, the contact pressure on the seal (seal pressure) is markedly reduced when stress is applied and if the seal taper is too small, wear occurs on the sealing surfaces due to an increase in sliding distance during build-up and breakage.

Based on the above results, the present inventors realized that if both the thickness of the flange and the length of the flange were increased (lengthened) as much as possible, the tightness with respect to external pressure can be greatly improved and that the compressive strength can be improved at the same time.

With respect to threads, the application of conventional techniques of US Patent Nos. 5,829,797 and 5,419,595 to the thread shape is best if only compressive strength is taken into account. However, these conventional techniques have the problems described above and are not suitable for mass production.

The present inventors realized that since the compression strength by the protrusion itself can be greatly improved by the shape of the flange described above, which has an increased flange thickness, even though ultra-high performance threads are not employed which They are not easily produced, the compressive strength of an overall joint can be significantly improved by employing high performance easily produced threads.

Based on this idea, the present inventors conducted the embodiment investigation with respect to the finding that the longer the flange length, the better the tightness with respect to external pressure. As a result, it was found that the sealing surfaces of the tip and the box are in contact only on the underside of the tip (the side closest to the threads), while an increase in the volume of the remaining portion that does not making contact of the flange from the end of the part in contact to the tip of the flange serves to increase the rigidity of the flange against external pressure.

In the embodiment of Figure 4, which does not represent the present invention, even if the length of the flange increases, since the non-contacting part of the flange from the end of the sealing surface to the tip of the flange is tapered, increasing the length of the flange does not significantly increase the volume of the non-contacting part of the flange and undesirably results in a decrease in the projecting surface area, thus decreasing the ability of the protruding part to withstand a compressive load. Therefore, the shape of the flange was changed to that shown in Figure 1.

In Figure 1, to increase the wall thickness of the flange surface as much as possible and increase the volume of the non-contacting part of the flange from the sealing surface to the end of the flange as much as possible, a part 15 (hereinafter referred to as the front part) of a flange 12 between a sealing part 13 and a projecting part 13 of a tip member 1 is made almost cylindrical and the outer surface of the front part 15 is made not to contact with a cash member 2.

By giving the flange such a shape, the wall thickness of the projecting surface and the wall thickness of the sealing surface are satisfactorily increased as much as possible within a limited wall thickness of the pipe. However, due to manufacturing errors, the pipe cross section is not a perfect circle but has variations or ovalities in the wall thickness. Therefore, a chamfer 16 may be formed on the inner edge of the end portion of the tip by an amount corresponding to the manufacturing tolerance of the tubing to ensure that the inner edge of the tip end will not protrude beyond a circle. of a predetermined diameter centered on the joint axis. In such a case, the inner edge of the box 2 may also have a corresponding chamfer 26 as will be described later.

On the tip, if the thread is provided as close as possible to the sealing surface, the stiffness of the tip flange against external pressure is further increased and therefore the tightness with respect to external pressure is also increased. .

This concept is quite different from that of the prior art described in US Patent No. 4,624,488 and US Patent No. 4,795,200, in that the tip end is used as a protruding surface and therefore therefore, the thickness of the flange can be greatly increased compared to the prior art in which the protruding surface is in a different location.

Furthermore, because the wall thickness of the front part can be increased as much as possible, the tightness with respect to external pressure can be greatly increased, simply by slightly increasing the axial length of the front part (referred to below as the length lead).

Additionally, the projecting surface can be enlarged, and the projecting part can achieve its maximum capacity to support a compressive load. Consequently, if the compressive load bearing capacity is somewhat greater than that of the reinforcing thread, the compressive strength of the threaded joint can be considerably improved by using the front flange.

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Brief description of the drawings

Figure 1 is a schematic diagram illustrating the shapes of a flange and the periphery of the flange of a threaded joint for steel pipes in accordance with the present invention.

Figures 2 (a) and 2 (b) are schematic explanatory views of a premium joint of ordinary coupling type for OCTG, in which Figure 2 (a) is an enlarged view of a part of Figure 2 (b) and Figure 2 (b) is an overall sectional view.

Figure 3 is a schematic view illustrating the shape and dimensions of a trapezoidal thread exemplified by an API reinforcing thread.

Figure 4 is a schematic explanatory view of the factors determining the shape of a flange used in basic studies in connection with the present invention.

Figure 5 is a schematic explanatory view of the factors defining the shapes of the flange and the periphery of the flange of a threaded joint for steel pipes according to the present invention.

Figure 6 is a schematic explanatory view of the shapes of the flange and the periphery of the flange of a threaded joint for steel pipes according to the present invention, in which a projecting surface is at an angle.

Figure 7 (a) is a schematic explanatory view showing positional relationship between the internal diameter between a box groove and a line extending from the valley of a female thread, and Figure 7 (b) is a schematic explanatory view showing the positional relationship between a sealing surface of a tip and a line extending from the valley of a male thread or a tangent to the male thread and the front.

Figure 8 is a schematic diagram illustrating another embodiment of the present invention, in which a second projection is provided on each end face of a box.

Figure 9 is a schematic diagram illustrating another embodiment of the present invention, in which a second seal is provided at one end of a box.

Figure 10 is a schematic diagram showing an example of an embodiment of the present invention, in which a gasket of the present invention is provided on a tip member, the end of which has been drawn.

Figure 11 is a schematic diagram illustrating an integral gasket.

Figure 12 is a schematic diagram illustrating the sequence in which a load is applied to a sample during FEM analysis of the examples.

Figure 13 is a schematic explanatory view of a conventional premium gasket used as a comparative example.

Figure 14 is a schematic diagram illustrating another embodiment in which the male thread comprises a running portion where the male thread valley envelope has a reduced taper compared to that of the thread taper.

Figure 15 is a schematic diagram of another embodiment having a toric-conical sealing surface on the box, cooperating with the tapered surface of the tip.

Description of the preferred embodiments

The present invention will now be described in greater detail with reference to the figures and with respect to preferred embodiments. In the figures, the same members or parts of a joint have the same reference numerals.

Figures 1 and 5 are schematic explanatory views of a threaded joint for steel pipes according to the present invention, comprising a tip member 1 and a box member 2 connected to each other by a threaded connection. The tip member 1 has a male thread 11, at least one sealing surface 13 and at least one projecting surface 14 formed on the end part of a pipe. The box member 2 has a female thread 21, at least one sealing surface 23, and at least one projecting surface 24 corresponding to (capable of engaging or contacting) the male thread 11, the sealing surface 13 and the projecting surface 14, respectively, of the tip member 1 formed on the end of a pipe. This gasket is a type of premium gasket.

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In such a threaded joint for steel pipes, the constitution is completed when the projecting surface 14 on the end face of the tip member 1 abuts against the corresponding projecting surface 24 of the box member 2.

In accordance with the present invention, the tip member has a front portion 15 that does not come into contact with the corresponding portion of the box member 2 (i.e. the portion of box 2 facing the front portion 15 of tip 1 ). Leading portion 15 is located between sealing surface 13 (which is located on the outer periphery of the tip member closer to the end of the tip member than the male thread of the tip member) and projecting surface 14 (which is located on the end face of the tip member). As a result, since the length of the flange 12 increases without a decrease in the protruding surface 14, the tightness of the gasket against external pressure can be significantly improved. Furthermore, by providing the male thread 11 in the vicinity (and, preferably, in the immediate vicinity) of the sealing surface 13 of the tip 1, the stiffness of the flange 12 against reductions in diameter, and the tightness against external pressure increases. increases further.

With reference to Figure 5, in a preferred embodiment of the present invention, it has been found that a threaded joint for steel pipes, having the configuration described above, can be improved by specifying its shape and dimensions.

In particular, the present inventors carried out studies on the ranges of dimensions and shapes of each part to obtain, in a more efficient way, the effects described above of the present invention.

For a normal premium joint, a compressive strength equivalent to approximately 20% of the elastic limit of the pipe body is required, although some wells require a compressive strength equivalent to more than 60% of the elastic limit.

A compressive load is applied not only to the protruding part, but also to the thread. If a thread having a higher compressive load bearing capacity is used, the load on the protruding part can be correspondingly reduced. However, if the thickness of the flange 41 (the thickness of the tip wall in the center of the region of contact with the seal 50) has a flange thickness ratio (ratio of the flange thickness to the wall thickness of the pipe body) of at least 25% and preferably at least 50%, sufficient compressive strength can be obtained by the flange alone (without the use of a thread having a high compressive load bearing capacity ).

The upper limit of the flange thickness ratio can be increased by about 85% (of the wall thickness of the pipe body), if the drawing described below is performed.

Since the tightness with respect to external pressure improves as the thickness of the sealing part and the front part of the flange increase, the angle 46 of the chamfer 16 provided on the internal surface of the end of the flange is preferably 9 to 76 degrees relative to the joint axis.

However, an abrupt change in the internal diameter of the pipe can cause turbulence and erosion, such that the internal diameter 48 of the box is adjusted to be almost equal to the internal diameter 49 of the tip flange, and a chamfer 26 having a chamfer angle 47 and a tip-like shape is provided on the inner periphery of the projecting surface of the case. Thus, the chamfer angle of the box part is preferably in the range of 104 to 171 degrees for the same reasons given with respect to the chamfer of the tip.

In this way, the change in the diameter of the gasket is minimized, whereby the occurrence of turbulence can be avoided and, at the same time, the projecting part of the box is reinforced. Therefore, the extent of a compressive load capable of being borne by the projecting parts is increased to improve the compressive strength of the joint.

The leading length 45 depends on the size of the pipe, although it is approximately 4mm to 20mm for the size of pipes used as OCTG (which has an outer diameter of approximately 50 to 550mm).

As stated above, the longer the front length is the better, but the effect of improving the tightness is saturated when the front length reaches a certain level, so a maximum front length of 20mm is suitable for a real product. .

Preferably, when the male and female threads are tapered threads, the male thread comprises, on the side of the thread sealing surface, a running-in portion that is formed for the first time during cutting of the thread to make the male thread. . In this running-in portion, the male thread valley envelope has reduced taper compared to that of the thread taper (see Figure 14). Such reduced tapering of the thread valley envelope in the running-in portion produces several benefits: ease of machining of the tip sealing surface without scratches and an increase in flange stiffness, thus being stronger under external pressure. . Preferably, as described below, the running portion of the male thread does not mesh with the female threads. More preferably, an envelope of the male thread valleys in the running portion is a cylindrical surface.

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The shape of the sealing surface of the tip or box can be:

(i) a tapered surface resulting from rotation about the joint axis of a straight line inclined relative to the axis, or (ii) a bulged surface resulting from rotation about the joint axis of a curve and, more particularly, a toric surface if the curve is a circular ring or (iii) a toric-conical surface resulting from the rotation around the joint axis of a combined line of the inclined straight line and the circular arc.

Preferably, the sealing surface of one of the tip and the case (for example, that of the tip) is a tapered surface, while the sealing surface of the other member (for example, that of the case) is a toric surface. or a toric-conical surface, the tapered portion of the toric-conical sealing surface being on the front side and having substantially the same taper as the tapered sealing surface (see Figure 15).

The cooperation of a tapered sealing surface and a toric-conical sealing surface has shown excellent stability of contact pressure (and thus tightness) for various conditions of use, for example, resulting in a charge cycle such as the one in Figure 12.

For toric-conical sealing surfaces domed, toric or with a toric part, the radius of curvature of the surface is preferably greater than 20mm and more preferably greater than 40mm.

For the reasons stated above, the angle of the sealing surface relative to the axis of the gasket, that is, the taper of the seal 44 is 5 to 25 degrees, and more preferably 10 to 20 degrees.

The existence of a taper of the seal implies, at the tip, a surface discontinuity between the sealing surface and the external surface of the front part.

The sealing surface requires a substantial contact region having an axial length of at least about 1 to 1.5 mm to ensure sealing properties.

However, if the sealing surface is too long, a sufficient wall thickness of the protruding part of the torque cannot be obtained and the cost required to finish the sealing surface increases, resulting in a reduction in productivity.

Therefore, the axial length 50 of the region of the sealing surface, where there is substantial contact, is 2 to 8 mm, and preferably 3 to 5 mm.

The projecting surface may be substantially perpendicular to the axis of the joint, as shown in Figure 5. However, according to the investigations described above, the tightness when a compressive load is applied improves if a hook angle is provided, as shown in Figure 6, although if the hook angle is too large, the stiffness of the protruding surface of the box is reduced, resulting in a decrease in compressive strength. Therefore, if a protrusion angle 43 is provided, it is preferable that it be 4 to 16 degrees with respect to a plane perpendicular to the axis of the joint.

Figure 6 is a schematic explanatory view of the shape of the flange and the periphery of the flange when a hook angle is provided on the projecting surface of a threaded joint for steel pipes in accordance with the present invention.

In these embodiments, the tip overhang is a single surface that extends seamlessly between an outer surface and an inner surface of the tip.

Interferences are present between the sealing surfaces and between the threads of the tip and the housing. If the sealing surfaces are too close to threadably engage the thread portions, the extent of substantial interference between the sealing surfaces is undesirably reduced due to interference between the threads.

However, as regards the tip, as described above, the presence of an additional thread in the part between the thread and the sealing surface is desirable to increase the stiffness of the tip against external pressure.

Thus, the structure of the tip is preferably such that the male thread extends to be as close as possible to the sealing surface. For this purpose, the sealing surface is preferably spaced from the male thread by a distance equal to at most one thread pitch to increase the stiffness of the flange. At the same time, a circumferential groove is provided in the case to prevent the thread parts near the sealing surfaces of the tip and the case from meshing with each other. If a running-in part is provided on the thread

ES 2 349 166 T3 male, the running part of the male thread corresponds to the part just mentioned of the male thread, which is prevented from meshing with the female thread by the groove 32.

If the axial length 51 of the circumferential groove 32 is small, the substantial interference between the sealing surfaces is reduced due to the interference of the thread. On the other hand, if the axial length 51 of the groove 32 is too large, the sealing capacity is reduced and the manufacturing costs are unnecessarily increased. A preferred range for the length 51 of groove 32 is 1.5 to 3.5 times the thread pitch.

Figure 7 (a) is a schematic explanatory view showing the positional relationship between the internal diameter of the groove of the box and a line 63 extending along the valley of the female thread, and Figure 7 (b) is a schematic explanatory view showing the positional relationship between the sealing surface of the tip and a line 62 extending along the root of the male thread or tangent 61 to the male thread and the front.

As shown in Figure 7 (a), the internal diameter of the circumferential groove of the box is preferably such that the internal surface of the groove is disposed radially outward from the extension line 63 of the valley of the female thread with respect to the axis. from the box. However, if the internal diameter of the groove increases excessively, the external diameter of the box also increases disadvantageously. Therefore, the internal diameter of the groove must be greater than the radial distance from the axis of the housing to the extension line 63 of the valley of the female thread only to the extent that fabrication can be easily accomplished.

As shown in Figure 7 (b), the positional relationship between the thread, the sealing surface and the front (the outer surface) of the tip is such that the sealing surface is arranged radially inward (as shown measured from the axis of the tip), both tangent 61 to the crest of the male thread and the apex of the tip (the front) as well as the extension line 62 of the valley of the male thread.

By employing such an arrangement, it is made difficult for the sealing surface to be damaged even though the tip end hits something during field work, and a reduction of the tightness can be prevented.

The gap between the drilling flanks and the angle of the drilling flank has a significant influence on the compressive strength of the threads. As previously described, the gap between the drill flanks is 30 pm to 180 pm for an API reinforcing thread. Assuming that the error occurrence rate for an API reinforcing thread has a normal distribution, the products that have a gap between the drill flanks of approximately 100 pm will be the most numerous. However, such a gap size does not produce a particularly high compressive strength.

By conducting investigations based mainly on the finite element method, the present inventors found that a high compressive strength is obtained if the gap between the drilling flanks is at most 90 pm.

However, if the piercing gap is too small, the pressure (called dopant pressure) of a lubricant called dopant that is coated on the threads during forming can rise to an abnormal level and adversely affect the performance of the joint. Additionally, if the drilling flanks come into contact with each other, as described in US Patent No. 5,829,797 described above, wear or variations in build-up torque can occur.

In this way, the effects of the present invention can be obtained by setting a gap between the perforation flanks, to be in the range of 10 pm to 150 pm, in view of the normal distribution of manufacturing errors. Preferably, the gap between the perforation flanks is 20 to 90 pm to obtain further improved effects.

The ability of the threads to withstand a compressive load increases as the angle of the drilling flank decreases. Conversely, as the angle of the drilling flank increases, the male thread and female thread slide along a slope and the ability of the threads to withstand a compressive load is reduced. The upper limit of the angle of the drilling flank, so that the ability to withstand a compressive load is not seriously reduced, is approximately 36 degrees.

On the other hand, if the angle of the drilling flank is too small, the damage due to wear of the thread cutting tool is aggravated, and it becomes very difficult to maintain manufacturing tolerances. Therefore, the lower limit is set to 3 degrees.

In the above description, an embodiment of the present invention was described in which a sealing surface and a protruding surface are provided on the flange at the end of a tip, but embodiments other than those described above are possible for steel pipes for special applications.

For example, as shown in Figures 8 and 9, when the present invention is applied to a drill pipe with an upset end or an OCTG with an extremely large wall thickness, an embodiment in which a second is provided is possible. series of projecting surfaces 33 and / or a second series of

ES 2 349 166 T3 sealing surfaces 34, with the tip 1 and the box 2 in combination with the structure of the lip of the tip 12, having the front according to the present invention.

Figure 8 is a schematic diagram illustrating an embodiment of the present invention having a second series of tip and housing projecting surfaces on the end face of the housing.

Figure 9 is a schematic diagram illustrating an embodiment of the present invention having a second series of tip and box sealing surfaces at the end of the box.

Figure 10 is a schematic diagram showing an example of an embodiment of the present invention, in which the gasket of the present invention is provided on a tip member having an end that has been drawn to reduce both the diameter of the tubing at the end so as to increase the wall thickness at the tip flange.

The present invention provides a sufficient tightness even though the tip end has not been drawn. However, if it is desired to obtain a very high level of torsional strength or compressive strength, for example, the wall thickness of the protruding surface can be increased by forming the threaded joint of the present invention over the end of a point that is has previously subjected to drawing (see Figure 10). In this case, however, the change in the internal diameter of the gasket should be properly selected so that turbulence does not occur.

Embodiments of the present invention have been described using a coupling type joint as shown in Figure 2 as an example. However, the present invention as described above can be applied to an integral joint as shown in Figure 11 (one in which the pipes are directly connected to each other without using a coupling but instead providing a tip. at one end of a pipe body and providing a box at the other end).

Figure 11 is a schematic diagram illustrating an integral joint in which a tip member 1 and a box member 2 provided at the respective end portions of the pipe bodies directly mesh with each other to form a threaded joint.

Examples

To clearly demonstrate the effects of the present invention, a numerical simulation analysis was performed using the elastoplastic finite element method for the samples shown in Table 1.

The samples shown in Table 1 were all OCTG coupling type threaded joints, such as those shown in Figure 2 for use with 14 cm (5-1 / 2 in.) Steel pipe, No. 20 (lb / ft) (139.7mm outer diameter and 9.17mm wall thickness). The steel material used for all samples was specified as P110 in the API standards.

Sample B was a comparative example in the form of a conventional premium gasket. It had a tip rim shape as shown in Figure 13. In Figure 13, reference numerals that are the same as previously used indicate the same members.

Samples C-O had basically the same gasket structure as for Sample A, except that the dimensions of the various parts were varied as shown in Table 1.

However, for sample J, the end of a pipe body having the size described above was upset by external overlap so that its outer diameter increased to 148.4 mm and, as shown in Figure 8, each second projection, consisting of a flat surface perpendicular to the joint axis and having a thickness corresponding to the amount in which the pipe body was upset, It is provided on the end face of the box and on the corresponding tip.

Sample L was another comparative example in which the protruding surfaces were provided on the end face of the box (instead of on the end face of the tip according to the present invention) and on the corresponding position of the tip.

The threaded shape as defined in Figure 3 will be the mass for all samples with respect to taper (1/18), thread height 74 (1.575mm), thread pitch (5.08mm), and the angle of the loading flank 71 (3 degrees). The axial gap 73 of the pierce flanks and the angle of the pierce flank 72 were varied and are shown in Table 1, along with the other gasket dimensions, including the flange thickness ratio (the flange thickness ratio 41 to the thickness of the pipe wall, where the thickness of the flange was measured at the center of the contact sealing part 50), the length of the front part (front length) 45, the taper of the sealing surfaces (seal taper) 44, the axial length of the seal contacting part (seal length 50), the boss angle 43, the chamfer angles of the tip and box bosses, 46 and 47, the presence or absence of the circumferential groove 32 and the distance between the thread and the sealing surface of the tip (in terms of thread pitch) (see Figures 5 and 6).

ES 2 349 166 T3

In finite element analysis, the material was molded as an elastoplastic material with an isotropic hardening having an elastic modulus of 210 GPa and a nominal yield strength in the form of a displacement performance of 0.2% of the yield point of 110 ksi (578 MPa).

The constitution was carried out for each sample, contacting the protruding surfaces of the tip and the box and then performing a 1.5 / 100 rotations of rotation.

In a first analysis, a compressive load (2852 kN) corresponding to 100% of the elastic limit of the pipe body was applied to each constitution sample, and the compressive strength was evaluated based on the proportion of the defined residual torque. as the value of the remaining maintained torque after removing the load (corresponding to the breaking torque, which is the connection-free torque after loading) divided by the constitution torque (the higher the ratio, the less loosening of the joint will occur. A value of at least 405 is considered necessary ).

In a second analysis, the loading sequence shown in Figure 12 and Table 2 was applied to each constitution sample. The tightness with respect to external pressure was evaluated by comparing the minimum values ​​of the mean seal contact pressure in the sequence (the larger the value, the better the tightness).

The results are shown in Table 3. From the results in Table 3, it can be seen that a threaded joint for steel pipes according to the present invention had a higher level of resistance of the remaining torque and maintained a higher seal contact pressure than any of the comparative gaskets and therefore had excellent compressive strength and tightness with respect to external pressure.

In this way, according to the present invention, a threaded joint for steel pipes is obtained having excellent compressive strength and tightness with respect to external pressure. At the same time, handling of the gasket in the field can be facilitated. Although the present invention has adequate performance even if the embossing processing is not performed, if embossing is performed to increase the wall thickness of the tip flange, not only compressive strength and tightness with respect to external pressure, but also torsional strength. For special pipes such as pipes with extremely large wall thickness, by additionally providing a second protrusion and / or metal-to-metal seal part, a high level of compressive strength can be obtained such as torsional strength and / or tightness with respect to the external pressure.

Although the present invention has been described with respect to preferred embodiments, they are merely illustrative and are not intended to limit the present invention. Those skilled in the art should understand that various modifications can be made to the embodiments described above without departing from the scope of the present invention, as set forth in the claims.

(Table goes to next page)

ES 2 349 166 T3

Table 1

<td>Overhang at tip end</td><td>ώ</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td>ω</td><td rowspan="13">*: The flange thickness ratio was the ratio of the flange thickness to the pipe wall thickness, where the flange thickness was the wall thickness of the tip at the center of the contact portion of the seal.</td>

<td>Drill flank angle [°]</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td><td> 45</td><td>OR</td><td> 45</td><td>OR</td><td> 45</td><td>OR</td><td>OR</td>

<td>Drilling flank gap [pm]</td><td> 09</td><td> 100</td><td> 09</td><td> 09</td><td> 09</td><td> 09</td><td> 09</td><td> 09</td><td> 09</td><td> 09</td><td> 160</td><td> 09</td><td> 160</td><td> 160</td><td> 09</td>

<td>y &lt;S&gt; O _, -, θ ^__m = Oj EO ro 4 = CO w tu 3 co ω co or. Q (D 2 (D -</td><td> 0,4</td><td></td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td> 0,4</td><td></td>

<td>Box groove</td><td>Apply</td><td>N / A</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td><td>N / A</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td><td>Apply</td>

<td>g&gt; £ -5 £ φ * &lt;€ 8 or</td><td> 165</td><td> 180</td><td> 165</td><td> 165</td><td> 165</td><td> 165</td><td> 174</td><td> 165</td><td> 165</td><td> 165</td><td> 174</td><td> 165</td><td> 165</td><td> 165</td><td> 165</td>

<td>5 -ra _ £ ra ω φ &lt;€ 8</td><td>IT</td><td>OR</td><td>IT</td><td>IT</td><td>IT</td><td>IT</td><td>CO</td><td>IT</td><td>IT</td><td>IT</td><td>CO</td><td>IT</td><td>IT</td><td>IT</td><td>IT</td>

<td>Boss angle [° 1</td><td>OR</td><td>IT</td><td>OR</td><td>OR</td><td>OR</td><td> 20</td><td>or</td><td>OR</td><td>OR</td><td>OR</td><td>or</td><td>OR</td><td>OR</td><td>OR</td><td>OR</td>

<td>Seal axial length [mm]</td><td>CO</td><td> 3,3</td><td>CO</td><td>CO</td><td>CO</td><td>CO</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>CO</td>

<td>Seal taper [° 1</td><td>CO</td><td> 20</td><td>CO</td><td> 30</td><td>CO</td><td>co</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>CO</td>

<td>Front length [mm]</td><td>CO</td><td>OR</td><td>C \ 1</td><td>CO</td><td>CO</td><td>co</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>co</td><td>CO</td><td>CO</td><td>CO</td><td>CO</td>

<td>Flange thickness ratio * [%]</td><td> 63</td><td> 46</td><td> 63</td><td> 63</td><td> 56</td><td> 63</td><td> 63</td><td> 63</td><td> 63</td><td> 63</td><td> 63</td><td> 23</td><td> 63</td><td> 63</td><td> 63</td>

<td>Sample</td><td> &lt;</td><td>co</td><td>OR</td><td>Q</td><td>LLl</td><td>or_</td><td> 0</td><td>T</td><td> —</td><td> “3</td><td></td><td> —1</td><td></td><td>z</td><td>OR</td>

ES 2 349 166 T3

TABLE 2

<td rowspan="2">Loading Stage</td><td rowspan="2">Conditions of Load</td><td colspan="2">Load size</td>

<td>Axial load</td><td>External pressure</td>

<td> 1</td><td>Constitution (1.5 / 100 turns from ledges)</td><td>0 kN</td><td>0 MPa</td>

<td> 2</td><td>PBYS compression 100%</td><td>-2852 kN</td><td>0 MPa</td>

<td> 3</td><td>100% PBYS compression + 100% API external pressure</td><td>-2852 kN</td><td>76.5 MPa</td>

<td> 4</td><td>API 100% external pressure (compression load removed)</td><td>0 kN</td><td>76.5 MPa</td>

<td> 5</td><td>PBYS voltage at 67% + API external pressure 100%</td><td>1911 kN</td><td>44.1 MPa</td>

<td> 6</td><td>PBYS voltage 100% (external withdrawal pressure)</td><td>2852 kN</td><td>0 MPa</td>

PBYS: Elastic Limit of the Pipe Body, that is, the elastic limit of the pipe body.

API: Collapse pressure (external crushing pressure) of the pipe body defined in the API 5C3 standard.

TABLE 3

<td>Sample</td><td>Proportion of maintained torque (%) obtained in the first analysis *</td><td>Minimum value of mean seal contact pressure (MPa) obtained in the second analysis</td>

<td>TO</td><td> 74</td><td> 301</td>

<td>B</td><td> 33</td><td> 23</td>

<td>C</td><td> 68</td><td> 151</td>

ES 2 349 166 T3

<td>Sample</td><td>Proportion of maintained torque (%) obtained in the first analysis *</td><td>Minimum value of mean seal contact pressure (MPa) obtained in the second analysis</td>

<td>D</td><td> 62</td><td> 226</td>

<td>AND</td><td> 59</td><td> 233</td>

<td>F</td><td> 49</td><td> 203</td>

<td>G</td><td> 61</td><td> 242</td>

<td>H</td><td> 71</td><td> 253</td>

<td> 1</td><td> 48</td><td> 287</td>

<td>J</td><td> 78</td><td> 315</td>

<td>K</td><td> 58</td><td> 236</td>

<td>L</td><td> 28</td><td> 36</td>

<td>M</td><td> 63</td><td> 211</td>

<td>N</td><td> 58</td><td> 233</td>

<td> 0</td><td> 43</td><td> 108</td>

<td colspan="3">* (torque maintained after load removal) / (torque constitution)</td>