JP2000081173A - Screw joint for oil well pipe - Google Patents

Abstract

PROBLEM TO BE SOLVED: To keep airtightness even after a load of high compression force is received and minimize the looseness under tensile and compressive axial force fluctuation by setting the screw interference quantity defined by the difference in pitch inner diameter between a male screw and a female screw. SOLUTION: The screw interference quantity defined by the difference in pitch inner diameter between a male screw and female screw geared in 1:1 in the fastening completion of a joint is set to a right value. The upper limit value of the screw interference quantity is set to 2 times the smallest value of values calculated extending over the whole length of the screw part by use of expressions I and II (wherein δ1, δ2 represents apparent interference quantities for radius portion (mm), σy represents the yield strength of a joint material (kg/mm2), E represents the Young's modulus (kgf/mm2), D represents the pin inside diameter (mm), and dp represents the pitch circle diameter (mm); La2=dp2-d2, Lb2=D2-dp2). The lower limit of the screw interference quantity is set to 5% of the upper limit of the screw interference quantity.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

TECHNICAL FIELD The present invention relates to oil, natural gas,
Oil and gas wells for exploration and production of steam and hot water, injection wells for injecting and disposing of industrial waste underground, and injection wells for secondary and tertiary recovery of oil, natural gas, etc. The present invention relates to a threaded joint for pipe connection. Hereinafter, these are referred to as “oil well pipes” and threaded joints for oil well pipes (
Or simply a joint) for convenience.

[0002]

2. Description of the Related Art Conventionally, threaded joints have been used for oil country tubular goods, and API standards are currently most frequently used.
(American Petroleum Institute standard) round threaded joints or buttressed threaded joints (hereinafter referred to as API joints). There has been a demand for a joint having high strength against breakage due to cracking and good airtightness.

FIG. 1 (a) shows a coupling type oil well pipe threaded joint for connecting a coupling 20 having two box portions 21 at both ends and a steel pipe 10 having pin portions 11 at both ends as an example. FIG. 2 is a cross-sectional view including a tube axis (hereinafter simply referred to as a cross-sectional view). FIG. 1 (b) shows the box portion 21 of the coupling 20.
FIG. 1 (c) is an enlarged view showing a cross section of FIG.
It is an expanded sectional view of 11.

As shown in FIG. 1A, a steel pipe as an oil well pipe
10, 10 'are screw-connected by a coupling 20,
Box sections 21 on both sides of coupling 20 and steel pipe
Pin portions 11 are provided at the tips of 10, 10 '. FIG.
(b) As shown in FIG.
The pin portion 11 is provided with a tapered male screw 12 as shown in FIG. 1 (c).

As described above, the coupling type threaded joint comprises a pin portion 11 having a tapered male thread 12 provided at an end of a steel pipe 10 and a tapered female thread provided inside a coupling 20.
The two steel pipes 10, 10 'are connected by screwing together a box part 21 with 22.

However, as described above, to meet the demand for a joint having high strength against tensile force due to its own weight and high airtightness,
A joint consisting only of a screw element consisting of a male screw 12 and a female screw 22 (eg an API joint) is not sufficient, and is shown in Figs. 1 (b) and (c).
As shown in FIG. 1, many of the pins 11 and the box 21 are provided with metal seals 13, 23 and torque shoulders 14, 24 at the same time.

In this metal seal portion, the amount of interference in the radial direction, that is, the outer diameter of the metal seal portion 13 of the pin portion is larger than the inner diameter of the metal seal portion 23 of the box portion, and this difference is called the amount of interference. When the joint is screwed in, the interference amount generates a surface pressure on the contact surfaces of the two metal seal portions, and it is expected that good airtightness is maintained by the surface pressure.

The provision of the torque shoulder portions 14 and 24 is because the pin portions 11 and the box portion 21 are brought into contact with each other, so that a high contact surface pressure that causes excessive plastic deformation is generated. This is because a sufficient screwing amount is ensured so as not to occur in the screw joint and the fastening of the threaded joint is ensured. In addition to the metal seal part, the screw part also has the same amount of interference as the metal seal part in order to secure its fastening and prevent loosening easily. Due to the regulation, the amount of interference of the screw portion is also limited to the safe range, and the occurrence of excessive stress in the box portion is suppressed.

Therefore, it can be said that such a coupling type screw joint has a structure suitable for use in the above-described corrosive environment. The angle of the torque shoulder portions 14, 24 with respect to the vertical plane is referred to as a shoulder angle (θ).

[0010] In recent years, the well depth has been deep,
Along with the deterioration of the well environment, on the other hand, the development technology of oil and gas wells has also advanced, and in recent years, the following severe requirements have been required for joints.

1) Withstand the tensile force in the axial direction due to the weight of the connected pipe. 2) Withstand the internal pressure of the internal fluid or the external pressure of the external fluid. 3) The pipes used, such as corrosion-resistant materials, tend to be expensive, so they can be used repeatedly. 4) Sufficient resistance to compression and torsion applied to pipes during the well development process. 5) Withstand internal and external pressure seals even when repeatedly receiving the above load.

In response to such demands, many proposals have been made regarding a joint having a metal seal portion and a torque shoulder portion in a fastened state, for example, as exemplified below.

(A) In the joint proposed in Japanese Patent Application Laid-Open No. 5-87275, the metal seal portion has a steep taper of 20 ° so that the metal seal portion is seized even in a relatively large diameter. And the effect of increasing the frequency of reuse is obtained. Also, the gap between the negative load surface flank angle and the insertion surface when fastened is set narrower than the API buttress screw, and has sufficient strength against simple tensile force, and at the time of tensile load after receiving compressive load Has a relatively high internal pressure resistance.

(B) In the example disclosed in Japanese Patent Application Laid-Open No. 9-119564, (i) restrictions are imposed on the flank angle of the screw load surface and the flank angle of the insertion surface. (Iii) Both the load surface and the insertion surface of the male screw and female screw are in contact with each other at the time of fastening and at the completion of fastening, and gaps are provided on both the top surface and the bottom surface. It is possible to provide a joint having excellent tensile performance, excellent airtightness even after compressive load, and high resistance to loosening torque.

(C) Japanese Patent Application No. 8-25 filed by the present applicant
In the example shown in No. 1163, in a joint provided with a seal forming portion and a shoulder forming portion at the pin tip and the back portion of the box, respectively, the form factor of the shape of the seal lip portion (i) the shoulder angle (°), ( ii) lip thickness ratio, (iii) lip length,
(iv) The seal length and (v) the seal taper are appropriately selected to provide a joint having excellent external pressure resistance.

(D) Examples disclosed in Japanese Patent Application Laid-Open No. Hei 9-119565 are as follows: (i) the shoulder angle is set to 5 ° to 20 °;
i) By selecting lip ratio (lip base thickness / lip outer diameter) / (inner pipe thickness / tube diameter) ≧ 0.52, it is possible to provide a joint having excellent external pressure resistance.

[0017]

SUMMARY OF THE INVENTION An object of the present invention is to provide a threaded joint for oil country tubular goods in which the joint performance of these conventional threaded joints for oil country tubular goods is further improved.

Specifically, the object of the present invention is, as illustrated in FIG. 2, as shown in FIG. As long as tension and compression occur on the inside and the section does not deform, it can be considered to be equivalent to the load of the axial force.)
Includes the entire inside of the ellipse with 100% equivalent stress (VME), and the external pressure area is within the API collapse pressure (considering axial force) 100% specified by API5C3, or the inside of the ellipse with 100% equivalent stress Von Mises. It is a joint that can maintain internal / external pressure leakage resistance and joint strength under compound load conditions including all smaller areas, and (ii) the above Von Mises equivalent stress of 100.
% Of the ellipse and the API5C3 collapse value ellipse, no matter how the load changes repeatedly, it is a joint that can maintain its leak resistance and joint strength, and can be reused many times. To provide.

As a measure of safety against repetitive fluctuating loads, the torque holding rate after repetitive fluctuating loads (the ratio of the holding torque after loading to the fastening torque itself before receiving fluctuating loads) is used as an index, with 60% as a reference. And

Regarding these points, in the prior art, for example, the above-mentioned example (A), the joint becomes a reusable joint having a large diameter, and has a high compression resistance and a high compression resistance of the VME stress ellipse.
The range up to about 60% can be covered, but under a higher compressive force, it is particularly difficult to seal after unloading, and the resistance to repeated loads is low.

Similarly, in the case of (B), by setting the dimensional restriction of the thread form, it has a sufficiently high tensile strength (100% or more of the tubular body) and also a high compressive resistance and a compressive load as in the conventional case. Although it also has airtightness after compression, it is still airtight after a compressive load of 50% tube strength and cannot be loaded up to 100% compression.

In the example (C), excellent external pressure resistance is obtained by restricting the shape and size factor of the lip portion and satisfying the related linear expressions, but the resistance to load fluctuation is not improved. Unexpectedly, for example, when subjected to a high compressive force, it is unknown whether or not it has sufficient external pressure resistance after unloading.

In the example of (D), excellent external pressure resistance is obtained by the limitation of the shoulder angle and the limitation of the lip ratio.
For example, as in the case of (C), it is unknown whether or not it has sufficient external pressure resistance after unloading when subjected to a high compressive force.

That is, both the examples (C) and (D) have a high external pressure resistance when the pin and the box are fastened, but for example, when unloaded when subjected to a high axial compression force, It is unknown whether the external pressure resistance can be maintained as it is. The same applies to the internal pressure resistance.

[0025]

Means for Solving the Problems In order to solve the above-mentioned problems, the present inventor has conducted a detailed analysis and study on the behavior when a threaded joint having a metal seal portion and a torque shoulder portion receives various loads. As a result, the following items were found.

(I) For a tensile force, a conventionally used trapezoidal screw with a negative load surface flank angle (an improved form of the API buttress screw) and a complete engagement length of the screw is equal to the pipe body. At least three times the wall thickness, and take the pin-side thread as long as possible without changing the taper until the root surface is erased on the outer surface of the pipe, and make it sufficiently compatible with the box-side screw to achieve a tensile strength higher than the strength of the pipe body. It can be applied that strength can be obtained. Of course, it is a prerequisite that the thickness and strength of the box side is sufficient, and depending on the purpose, the strength of the box side may need to be reduced due to the outer diameter restrictions of the joint. does not change.

(Ii) With respect to the compressive force, in the case of a screw having a gap in the axial direction, which is conventionally used, when a compressive force close to the strength of the tube is received, the torque shoulder portion which receives the load first is naturally larger than the tube. In addition, the contact area is small and the load cannot be supported, yielding and causing plastic deformation. The magnitude of this plastic deformation in the axial direction depends on the axial gap of the screw portion, and this is inevitable as long as there is a gap. This deformation of the torque shoulder results in an axial displacement of the loose side between the pin side and the box side of the metal seal part, especially in the case of a steep seal taper that is selected to be reusable many times. The looseness is remarkable, and has a significant effect on the leak resistance under tension after compression load.

Therefore, in order to prevent the compressive force from directly affecting the metal seal portion, it is necessary that the screw insertion surface always comes into contact from the middle of the fastening to after the fastening, so that a certain surface pressure is maintained. . In order to withstand a high compressive force equivalent to the strength of the tubular body, it is conceivable that at least a screw length necessary for a tensile axial force is necessary. In this case, Japanese Patent Application Laid-Open
The contents discussed in -119564 can be used.

(Iii) With respect to the internal pressure resistance (airtightness), sufficient airtightness can be obtained by providing a metal seal portion at the tip of the pin so as to have an appropriate amount of positive interference between the pin side and the box side. Is known to be obtained, and can be generally solved by adopting this.

However, since the presence of the interference amount of the screw portion exerts an action in the direction of separating the pin side and the box side of the metal seal portion, the interference amount of the metal seal portion is not larger than the interference amount of the screw portion. Not be.

(Iv) In order to ensure the external pressure resistance, a method of securing a lip ratio of 52% or more, which was studied and concluded in Japanese Patent Application Laid-Open No. 9-119565, or a method disclosed in Japanese Patent Application No. 8-251163. In consideration of the constrained form factors and dimensional constraints of the seal lip, and by giving a linear functional correlation found in experiments and analysis, the seal resistance against external pressure is calculated by API5C3. It can be understood that the pressure or more is secured.

(V) The above (i), (ii), (iii), and (iv) are all the results of studying a single load or a static load. (Production / injection work) Sometimes, various fluctuating loads are applied, and the tendency is remarkable due to the recent progress of development technology. Therefore, the performance of the joint can be trusted only when the performances (i) to (iv) described above can be maintained even with respect to a fluctuating load.

Therefore, in order to secure the stability against the fluctuating load, it is necessary that the screw portion does not loosen even when receiving a large change in the axial force, in other words, it is necessary that the torque applied at the time of fastening is not lost.

The pipe on which the threaded joint is machined is usually a steel pipe. Therefore, when the loads opposite to each other in tension and compression are applied repeatedly, and the large load near both the yield points is repeatedly applied, the properties and performance of the material itself are changed. It is an obvious fact that it is changing and cannot maintain its original properties and performance.

On the other hand, casing pipes (structural pipes constituting wells) and tubing pipes (pipes for producing underground fluids) used in oil wells and gas wells, etc., are subjected to a long period of use, and the pipes are more likely to yield. The number of times of tension / compression load is not so large, is about several times, and is in a range that can sufficiently withstand the tube itself.

Therefore, the same applies to the fluctuating load considered in such a joint portion of the pipe. It can be said that it is sufficient to withstand several repetitive loads for a fluctuating load as high as near the yield point of the pipe.

In order to secure the stability against the above-mentioned fluctuating load, it is necessary to add not only the above-mentioned (i), (ii), (iii) and (iv) but also the following conditions for the screw length. It turned out that there is.

Completely engaged screw length S> Three times or more the wall thickness of the pipe ・ ・ ・ When the wall thickness / outer diameter ratio is 0.096 or more Same as above S> 4 times or more the wall thickness of the pipe ・ ・ ・When the thickness / outer diameter ratio of the tube is 0.084 or more Same as above S> 5 times or more the wall thickness of the tube ・ ・ ・ Thickness / outer diameter ratio of the tube, other than the above, but outside the tube / thickness When the diameter ratio is 0.052 or less, the inner surface of the lip is reduced to the inner surface of the pipe as much as possible by increasing the thickness of the lip. In this specification, the ratio of (wall thickness of pipe) / (outer diameter) is simply referred to as the wall thickness / outer diameter ratio or pipe thickness to pipe outer diameter ratio.

Here, the present invention is based on the above findings, and the gist is as follows. (1) A thread portion is formed of a taper screw having a substantially trapezoidal shape, and includes a pin portion having a male thread and a box portion having a female thread, and a threadless portion for forming a metal seal portion provided on the pin portion, and a box portion having The metal seal portion is formed by the provided screwless portion for forming the metal seal portion, and the screwless portion for forming the torque shoulder portion provided at the tip of the pin portion and the screw for forming the torque shoulder portion provided on the box portion. A threaded joint for oil country tubular goods having a torque shoulder portion by abutting against a non-threaded part, characterized by satisfying the following items.

(A) The flank angle of the screw load surface is -20 ° C or more and less than 0 ° C, and the flank angle of the insertion surface is more than 25 ° and 45 ° or less.
Furthermore, there is a positive screw interference amount between the male screw and the female screw,
During the fastening of the joint and at the completion of the fastening, the load surface and the insertion surface of the male screw and the female screw are in contact with each other, and a gap is provided between the top surface and the bottom surface of the screw.

(B) As the screw interference amount (δ), of the values calculated for the total length of the screw portion using the following equation (1) or (2), twice the smallest value is set as the upper limit value. 5% of that as the lower limit.

[0042]

(Equation 1)

Here, δ: Radial thread interference amount (mm) σy: Yield strength of joint material (kgf / mm ² ) E: Young's modulus of joint material (kgf / mm ² ) D: Box outer diameter (mm) d: Pin inner diameter (mm) dp: Screw pitch circle diameter (mm) La ² : (dp ² -d ² ) Lb ² : (D ² -dp ² ) However, the amount of screw interference is based on the amount of metal seal interference. Small (screw interference <metal seal interference).

(C) The engagement length of the complete screw between the male screw of the pin portion and the female screw of the box portion is (a) at least three times the wall thickness of the pipe body at the time of fastening: the pipe wall thickness to pipe outer diameter ratio is 0.096 or more. (B) 4 times or more: と き When it is 0.084 or more and less than 0.096, (c) 5 times or more: It is other than the above (a) and (b). However, if the ratio of the pipe wall thickness to the pipe outer diameter is 0.052 or less, the inner diameter of the seal lip of the pin portion is set to the pipe inner diameter or less as much as possible.

(D) The shoulder angle of the torque shoulder is θ: 5 to 20 °, and the following formula between the thickness (S1) of the base of the lip of the seal lip and the wall thickness (Wt) of the pipe is obtained. The defined lip ratio is 0.52 or more. (S1 / DB) / (Wt / OD) ≥ 0.52 where DB: outer diameter at the base of the lip OD: pipe outer diameter However, when Wt / OD is small and Wt is thin, before pipe end swage etc. The rib ratio determined above satisfies ≧ 0.52 by processing.

Or (e) Factors determining the shape of the threadless portion at the tip of the pin portion and the threadless portion at the back of the box portion, lip length: We (mm):
The seal length: S (mm), the seal taper: Ts, the shoulder angle: θ (°), and the lip ratio: R are in the following ranges, and the function f, which is a first-order polynomial, is f> 1.
2mm 6mm ≦ We (mm) ≦ 30mm, 3mm ≦ S (mm) ≦ 10mm 1/16 ≦ Ts ≦ 1, 0 ° ≦ θ (°) ≦ 20 °, 0.25 ≦ R ≦
0.75 where f = −3.26 × 10 ^-1 + 3.19 × 10 ^-2 (1 / °) × θ (°) +1.4
3 × R−4.67 × 10 ⁻⁴ (1 / mm) × We (mm) + 8.39 × 10 ⁻² (1 / mm)
× S (mm) −6.22 × 10 ⁻¹ × Ts R = ｛(wall thickness at the root of the pin side lip) / (outer diameter of the root of the lip)｝ / ｛(wall thickness of the pipe body) / (tube body This is a threaded joint for oil country tubular goods, characterized by (outer diameter)｝.

[0047]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Next, the reasons for limiting each of the above constituent elements (a) to (e) according to the present invention will be described in detail. According to the configuration requirements (a) and (b), the pin side and the box side are in contact with both the screw load surface and the insertion surface from the middle of the screw tightening to the end of the screw tightening. This is a mechanism for loading from the beginning of the load. Therefore, an excessive load is not applied to the shoulder portion having the metal seal portion, and the deformation or deformation of the shoulder portion, and the destruction or deterioration of the sealing property during the tensile load after the compressive load considered to be caused by the deformation can be prevented.

Here, the reason why the angle of the flank of the insertion surface is limited to 45 ° or less is that when a compressive force is applied, the force is shared by the screw surface in the radial direction of the pipe, and the action of expanding the box portion works. In particular, if the angle exceeds 45 °, the effect is considered to be greater than the compressive force. The reason for limiting to more than 25 ° is to prevent seizure. The flank angle of the load surface of the screw is -20 ° to secure the tensile performance of the joint.
Above was less than 0 °. Preferably -10 ° or more, -3 °
It is as follows.

In the present invention, the screw interference amount defined by the difference between the pitch circle diameters of the male screw and the female screw that mesh with each other in a one-to-one manner when the joint is completed is set to a positive value, that is, the predetermined screw interference amount The reason is that a certain amount of torque is generated in the threaded portion even during the fastening of the joint to increase the fastening torque.

The upper limit value of the thread interference amount of the joint is as described in the above (1)
, (2) is twice the smallest value obtained from the expression. Here, the smallest value is the smallest value calculated by the formulas (1) and (2) over the entire length of the threaded portion because the pitch circle diameter changes linearly in the case of a tapered screw. In general, the thickness decreases at the tip of the pin portion and the box portion, and thus the value is calculated based on the specifications at the respective tip portions.

On the other hand, if the screw interference amount is too small, the contact surface pressure of the screw generated at the time of completion of the fastening becomes small, so that an appropriate torque cannot be generated in the screw portion. Therefore, the lower limit of the screw interference amount is set to 5% of the upper limit of the screw interference amount calculated from the above equations (1) and (2).

The reason why both the load surface and the insertion surface of the male screw and the female screw are brought into contact with each other during the joint fastening and at the completion of the fastening is, in part, to ensure the compression performance of the joint.
That is, even if a compressive force is applied to the joint, the male screw and the female screw do not shift from each other in the axial direction, and the compressive force is shared by the screw portions. Thereby, the effect of the insertion surface flank angle defined as described above can be sufficiently exerted.

Another reason is that the load surface and the insertion surface are brought into contact with each other even during the fastening, so that the contact surface pressure on the screw surface is appropriately increased, and the fastening torque at the time of completion of the fastening is maintained high. This is to prevent loosening.

(C) According to the present invention, since the pin portion and the box portion are in contact with both the load surface and the insertion surface of the screw,
The fastening torque is unlikely to be loosened, and has a tendency to be able to hold the fastening torque at a high rate because the torque shoulder portion of the seal lip portion is not deformed even under a high compression load. However, in order to secure the sealing performance during the tensile load after the compression load is removed from the actual threaded joint (in other words, to ensure the stability of the threaded joint performance), a certain length of screw is required. Yes, it is easy to imagine that it is more than three times the pipe wall thickness from the example of tensile load, but in the case of tension, the flank angle of the screw load surface is usually about + 3 ° to -10 °. It can be said that even if a tensile force is applied, the force exerts in the radial direction to expand the pin portion and the box portion, so that the effect of the action is small.
Further, it is preferable that the angle of the insertion surface, which is the surface on which the compressive force is borne, is set to a smaller angle standing on the tube axis. This is because when the compressive force is received, the force is dispersed on the insertion surface and the effect of expanding the box part is reduced, the hoop stress generated in the box part can be reduced, and the function of pushing the screw part of the pin part in the axial direction is also reduced. This is because the resistance to a large repetitive compression load increases.

However, the maximum value of the tolerance of the thread width in thread processing is required to be a minimum of 0.06 mm in consideration of mass production at a factory. In order to maintain the mechanism in contact with the box portion, it is necessary to provide a gap between the screw top surface and the bottom surface to absorb the clearance due to the processing tolerance in the screw width direction. However, increasing the clearance between the top and bottom surfaces means that the effective meshing width of the screw is constant, so that the height of the screw thread will be extra high, and the screw length will be long. This is disadvantageous in cost. Therefore, the gap between the top and bottom surfaces should be about 0.1 mm. If the angle of the load surface is -3 °, the flank angle θ of the insertion surface is
33.12 ° or more is required.

[0056]

(Equation 2)

Therefore, in the case of compression resistance, the required screw length is longer than the tensile resistance. A larger force acts in the radial direction than when a tensile load is applied. Moreover, this force becomes remarkable on the base side of the pin-side screw and on the side of the box-side screw close to the metal seal portion. If a high compressive force close to the tube strength is applied, a correspondingly large force acts on a narrow portion. Therefore, it is not difficult to imagine that a change occurs in the contact state of the flank surface in that portion, and the fastening torque also becomes small in that portion.
That is, if a large compressive load is applied several times, it is considered that the fastening torque gradually disappears from both ends of the screw connection portion between the pin portion and the box portion.

Therefore, the screw length required in the case of a compressive load resistance must take into account the length of the portion where the above-mentioned fastening relaxes under a high compressive load. As a result, a longer screw length is required. Is required.

In this case, it is considered that the likelihood of the thread portion being deformed by the radial force generated by the compression force depends on the rigidity of the annular body at that portion. When the diameter ratio is R≥0.096, the required full thread engagement length L≥3 times the pipe wall thickness When the wall thickness / outer diameter ratio of the pipe is 0.096> R≥0.084, the required full thread engagement length L≥ 4 times the pipe wall thickness Pipe wall thickness / pipe outer diameter ratio: 0.084> R ≧ 0.052 Required complete screw engagement length L ≧ 5 times the pipe wall thickness Pipe wall thickness / pipe outer diameter ratio: In the case of R <0.052, the required full thread engagement length L ≧ 5 times the pipe wall thickness However, in this case, the inner diameter of the seal lip of the pin part is lowered to the pipe body reference inner diameter, and (internal flush state) compression It has been found that it is necessary to support a part of the force with the torque shoulder surface to reduce the load on the screw portion.

Regarding the remaining external pressure resistance among the joint performance, as long as the external pressure sealing property is ensured even under a simple load condition, as long as the stability of the threaded joint is ensured by the above constitutional requirements (a) to (c). Not lost.

Therefore, it is a sufficient condition for maintaining the external pressure resistance that the above constitutional requirements (d) or (e) are satisfied. Thus, the object of the present invention is (a) to (d) or (a)
This can be achieved by satisfying (c) and (e).

FIG. 3 is an explanatory view of the dimensional relationship of the pin portion.
The following is a description in connection with the description of the present invention below. Pin part 11 is tube outer diameter (OD), tube inner diameter (ID), wall thickness
(Wt), the seal lip portion 17 is simply called a lip portion,
The length (We) is called the lip length.
(DB), root thickness (S1), and root inner diameter (Dy).
The length of the metal seal portion 13 is called a seal length (S), and its taper is called a seal taper (Ts).

Here, the lip ratio in the constituent requirement (d) is
The reason for setting it to 0.52 or more is to reduce the deformation of the lip portion of the pin portion with respect to the minimum collapse pressure specified by the API and to give the seal lip portion the rigidity necessary to maintain the sealing performance of the metal seal portion. Of course the shoulder angle (θ) is 5
The lip thickness alone is not always able to maintain such excellent sealing properties because the lip thickness is set to about 20 ° and a predetermined external pressure resistance function is provided. In other words, if the seal lip portion of the pin portion is deformed to some extent, the sealing mechanism of the metal seal portion is broken, and the external pressure seal resistance cannot be maintained.
The upper limit of the lip ratio is not particularly limited, but is preferably 55.
~ 65%. However, when Wt / OD is small and Wt is thin, if the rib ratio determined above does not satisfy ≧ 0.52, ≧ 0.52 is satisfied by pre-processing such as pipe end swage.

The reasons for limitation in the constituent requirements (e) are as follows. The range of possible values of the shoulder angle θ is set to 0 ° or more and 20 ° or less for the following reason. That is, the upper limit is set to 20 °, when the shoulder angle exceeds 20 °, excessive plastic deformation occurs particularly at the innermost shoulder of the box-side torque shoulder portion, and the torque limiting function is impaired and cannot be used. Because. The lower limit is set to 0 ° because when the shoulder angle is less than 0 ° (that is, the angle to the opposite side with respect to the cross section perpendicular to the tube axis = a> 0), the pin-side metal seal forming surface is boxed. This is because it acts in the direction of separating from the metal seal portion forming surface on the side, so that the sealing surface pressure of the metal seal portion decreases, and as a result, sufficient airtight performance cannot be obtained.

The reason why the lip ratio R is not less than 0.25 and not more than 0.75 is as follows. That is, the upper limit is set to 0.75 because, in practical use, it is not necessary to consider a lip thickness ratio exceeding 0.75. On the other hand, the lower limit is set to 0.25 because if the thickness of the lip at the tip of the pin is thinner than this, the rigidity of the lip will be too small, and the lip at the tip of the pin will not be excessively tight even if it is slightly overtightened. This is because so-called over-torque performance is deteriorated in that plastic deformation occurs and airtightness performance is deteriorated.

The range of possible values of the lip length We is set to 6 mm or more and 30 mm or less for the following reason. That is,
The gap (space) between the pin and the box that is formed in contact with the [parallel part without screw], which is the distance from the end of the metal seal part forming part to the end of the screw, is called [screw run-out groove]. However, the lip length is represented by (lip length) = (metal seal length) + [thread runout groove = parallel part without screw]. The thread run-out groove serves as an escape space for cutting chips of a threading tool called a chaser during threading on the box side. Therefore, the reason why the lower limit of the lip length is set to 6 mm is to secure a screw runout groove of a minimum required size. On the other hand, the upper limit is set to 30 mm, because even if the thread run-out groove length is unnecessarily increased, the airtightness, tensile resistance and compression resistance of the joint are hardly improved, and conversely the material cost is increased. is there.

The range of possible values of the seal length S is set to 3 mm or more and 10 mm or less for the following reason. That is, the reason why the upper limit of the seal length is set to 10 mm is that if the seal length is long, the contact surface of the metal seal portion forming portion becomes large, the contact surface pressure becomes small, and the airtightness is reduced. Further, if the fitting margin of the metal seal portion forming portion is increased in order to obtain a contact surface pressure enough to secure the airtight performance, seizure will occur on the metal seal portion forming surface.

On the other hand, the lower limit of the seal length is set to 3 mm for the following reason. That is, if the seal length is too short,
This is because the contact form of the metal seal portion forming surface becomes close to the line contact, and the contact pressure becomes too high, causing seizure or excessive plastic deformation of the metal seal portion forming surface.

Finally, the range of possible values of the seal taper Ts is set to 1/16 or more and 1 or less for the following reason. That is, when the upper limit of the seal taper is set to 1, the reduction rate of the contact surface pressure of the metal seal portion forming surface when a tensile force acts in the axial direction on the pipe body including the joint to be fastened if the upper limit is larger than this. This is because it is large, and the airtight performance is greatly reduced. On the other hand, the lower limit of the seal taper is 1/16,
If it is smaller than 1/16, the distance that the metal seal portion forming surface rotates while contacting, that is, the sliding distance becomes long, which causes seizure on the metal seal portion forming surface.

The function f shown in the inequality (d) of the present invention is to produce a coupling type oil country tubular good threaded joint having lip portions of various shapes within the range of the shape of the internal seal, It was determined by conducting a substantive test and analyzing the results.

[0071]

DESCRIPTION OF THE PREFERRED EMBODIMENTS The coupling type joint shown in FIG.
At this time, other specifications of the threaded joint are as follows.

Pipe outer diameter 177.8 mm Wall thickness Table 1 Screw form Load surface flank angle -3 ° Insertion surface flank angle 35 ° Screw height 1.575 mm Screw pitch 5.08 mm (5 ridges / inch) Top gap 0.10 mm Bottom gap 0.10 mm Screw type Double-sided contact screw between screw load surface and insertion surface Screw interference 0.20mm Metal seal interference 0.60mm Tube and coupling material 13Cr steel (yield strength 56kgf / mm ² ) (A) and (b) are satisfied,
In addition, the sample has a constant lip ratio (R) to satisfy (d). Note that the thread taper is determined uniformly for each sample as described above. In Example A, tube end swaging was performed.

Rip rate: R = {1/2 (DB-Dy) / DB} / (Wt / OD) The outer diameter of the box is 102% of the strength of the dangerous cross section of the box at each dimension. Set up and prototype.

FIG. 3 is a schematic explanatory view showing the dimensions of the joint used in this example. These threaded joints are subjected to repeated fastening tests, (tensile / compressive + internal pressure), (tensile / compressive + external pressure) tests,
A tensile internal pressure test after compression and unloading and a screw return test under repeated tension / compression were performed to evaluate the performance.

Here, in the repeated fastening test, lubricating grease was applied to the thread portion, and fastening and release of the joint were repeated until seizure occurred in the thread portion or the seal portion. The maximum number of repetitions was set to 10.

The (tensile / compression + internal pressure) and (tensile / compression + external pressure) tests were performed on a stress ellipse equivalent to 95% Vom Mises, while maintaining the internal pressure from 95% tension to 95% compression of the tube strength. After reciprocating load, API collapse pressure specified by API5C3 100%
Again, a reciprocating load was applied from 85% to 100% compression on the ellipse, and the presence / absence of internal / external pressure wetting under each combined load condition was examined. In addition, in the tensile + internal pressure test after unloading, after applying a compressive force of 95% tube strength in advance and unloading,
Apply 95% tension and increase the internal pressure so that the internal pressure becomes VME 95%, reduce the tensile force, increase the internal pressure, and change it with VME 95%,
The test of lowering the axial force to 95% CEYP and lowering the internal pressure afterwards was repeated three times.

In the unscrewing test under repeated tension / compression, a constant reverse torque equivalent to 60% of the fastening torque is applied to a joint fastened with a constant torque, and the tensile force and the compressive force are changed while the joint is maintained. Alternation was performed while increasing the pitch at a pitch of 70% to 5% of the body strength, and it was examined when the loosening of the screw occurred.

FIG. 4 shows (tensile / compression + internal pressure) · (tensile / compression)
It is a schematic diagram which illustrates the load application condition of a test (compression + external pressure). (Compression / tensile + internal pressure) Test: 0 → 1 → 2 → 3 → 4 → 5 →
6 → 7 → 0 → 7 → 6 → 5 → 4 → 3 → 2 → 1 → 0 (compression / tensile + external pressure) Test: 0 → a → b → c → d → e →
0 → ho → ni → ha → lo → i → 0 → i → he → to → ji → ri → 0
→ Re → H → G → F → A → 0 The above results are shown in Table 2. As is clear from the results shown in the table, the results of the present invention were quite good, but in the comparative examples, a failure occurred in any of the tests.

For example, in the joint H, leakage occurs due to the tensile internal pressure after the first compression, and looseness occurs after a 70% compression load in the torsion test. The other joints I to M have almost the same results, although there are some differences. On the other hand, the joint on the side of the present invention was able to show that in the (tensile / compression + internal pressure) and (tensile / compression + external pressure) tests, almost all areas of the stress ellipse equivalent to VME and the API collapse pressure ellipse were satisfied. To
(Although each joint of the comparative example also has a satisfactory result in this range, no leakage or the like occurs even in the tensile internal pressure test after a 95% simple compression load.

On the other hand, even in the loosening test, it was shown that there was no loosening up to a high axial force and the stability against axial force was high. Therefore, it can be said that it has sufficient stability against a fluctuating load.

Next, instead of the component requirement (d), the component requirement (e)
When the wall thickness was changed in the same manner as above for the joint satisfying the above, similar results were obtained, and the effectiveness of the present invention in this respect was confirmed.

[0082]

[Table 1]

[0083]

[Table 2]

[0084]

The threaded joint for oil country tubular goods according to the present invention is Vom
The entire range of internal pressure and axial force expressed by the Mises equivalent stress ellipse
API5C3 satisfies the entire range of external pressure and axial force expressed by the equation of the pipe collapse as well as the fluctuation of the axial force, especially 95% of the pipe strength
Even after receiving the load of high compression force, it keeps airtight,
In addition, it has a high resistance to loosening torque, which is difficult to loosen even under fluctuations in tension and compression axial force, so that it can be used satisfactorily under well development conditions that will become even more severe in the future.

[Brief description of the drawings]

1 (a) is an axial sectional view showing a basic configuration of a coupling type oil country tubular good threaded joint, FIG. 1 (b) is an enlarged sectional view of a box portion, and FIG. It is an expanded sectional view of a pin part.

FIG. 2 is a graph illustrating the yield region of Mises and the 100% collapse pressure region of API5C3.

FIG. 3 is a schematic explanatory view of a joint shape used in the embodiment.

FIG. 4 is an explanatory diagram of a procedure of an experiment in an example.

Claims (2)

[Claims] 1. A threadless portion for forming a metal seal portion provided on a pin portion, comprising a pin portion having an external thread and a box portion having an internal thread, each of which comprises a taper screw having a substantially trapezoidal thread shape. The metal seal part is composed of the metal seal part provided with the screwless part for forming the metal seal part, and the screwless part for forming the torque shoulder part provided at the tip of the pin part and the torque shoulder part provided for the box part A threaded joint for an oil country tubular good comprising a torque shoulder portion by abutting against a threadless part of the above, wherein the following items are satisfied. (B) The screw flank angle is -20 ° C or more and less than 0 ° C, the insertion surface flank angle is more than 25 ° and less than 45 °, and the screw interference defined by the difference between the pitch inner diameters of the male screw and the female screw. The amount must be a positive value, the load surface and the insertion surface of the male and female threads come into contact with each other during and during the fastening of the joint, and a gap must be provided on both the top and bottom surfaces of the screw. (B) The screw interference amount is
Of the values calculated for the total length of the threaded part using the formula (1) or (2), twice the smallest value shall be the upper limit, and 5% of the upper limit shall be the lower limit. (Equation 1) Δ ₁ , δ ₂ : apparent screw interference amount by radius (mm), σy: yield strength of joint material (kgf / mm ² ) E: Young's modulus of joint material (kgf / mm ² ), D: outside box Diameter (mm) d: Pin inner diameter (mm), dp: Pitch circle diameter of screw (mm) La ² : (dp ² -d ² ) Lb ² : (D ² -dp ² ) However, screw interference amount is metal seal Smaller than the interference amount of the part. (C) When the engagement length of the complete screw between the male screw of the pin part and the female screw of the box part is tightened, (a) 3 times or more of the pipe body wall thickness: When the pipe wall thickness to pipe outside diameter ratio is 0.096 or more, b) 4 times or more: と き When it is 0.084 or more and less than 0.096, (c) 5 times or more: When it is other than the above (a) and (b). However, if the ratio of the pipe wall thickness to the pipe outer diameter is 0.052 or less, the inner diameter of the seal lip of the pin portion is set to the pipe inner diameter or less as much as possible. (D) The shoulder angle of the torque shoulder portion is θ: 5 to 20 °, and the lip ratio defined by the following equation between the root thickness (S1) of the lip portion of the seal lip portion and the pipe wall thickness (Wt). Is set to 0.52 or more. (S1 / DB) / (Wt / OD) ≥ 0.52 where DB: outer diameter at the base of the lip OD: pipe outer diameter However, when Wt / OD is small and Wt is thin, before pipe end swage etc. The rib ratio determined above satisfies ≧ 0.52 by processing. 2. A screwless portion for forming a metal seal portion provided on a pin portion, comprising: a pin portion having an external thread and a box portion having an internal thread, each of which comprises a taper screw having a substantially trapezoidal thread shape. Forming a metal seal portion with a screwless portion for forming a metal seal portion provided in the portion, and forming a screwless portion for forming a torque shoulder portion provided at the tip of the pin portion and forming a torque shoulder portion provided in the box portion A threaded joint for oil country tubular goods having a torque shoulder portion by abutting against a threadless part for oil country tubular goods, which satisfies the following items. (A) The flank angle of the load surface of the screw is -20 ° C or more and less than 0 ° C, the flank angle of the insertion surface is more than 25 ° and 45 ° or less, and there is a positive screw interference between the male screw and the female screw. During the fastening and at the time of completion of the fastening, the load surface and the insertion surface of the male screw and the female screw are in contact with each other, and a gap is provided between the top surface and the bottom surface of the screw. (B) Among the values calculated for the total length of the screw part using the following formula (1) or (2) as the screw interference amount, the smallest value, 2
Double is the upper limit, and 5% of the upper limit is the lower limit. (Equation 1) Here, δ: Radial thread interference (mm) σy: Yield strength of joint material (kgf / mm ² ) E: Young's modulus of joint material (kgf / mm ² ) D: Box outer diameter (mm) d: Pin inside diameter (mm) dp: pitch diameter of thread ^{(mm) La 2: (dp} 2 -d 2) Lb 2: (D 2 -dp 2) However, thread interference is less than the amount of interference of the metal seal portion. (C) When the engagement length of the complete screw between the male screw of the pin part and the female screw of the box part is tightened, (a) 3 times or more of the pipe body wall thickness: When the pipe wall thickness to pipe outside diameter ratio is 0.096 or more, b) 4 times or more: と き When it is 0.084 or more and less than 0.096, (c) 5 times or more: When it is other than the above (a) and (b). However, if the ratio of the wall thickness to the outer diameter is 0.052 or less, the inner diameter of the lip of the pin shall be less than the inner diameter of the pipe as far as possible. (E) Factors that determine the shape of the threadless part at the tip of the pin and the threadless part at the back of the box, lip length: We (mm):
The seal length: S (mm), the seal taper: Ts, the shoulder angle: θ (°), the lip ratio: R are in the following ranges, and the function f, which is a first-order polynomial, is f> 1.
Satisfy 2. 6mm ≦ We (mm) ≦ 30mm, 3mm ≦ S (mm) ≦ 10mm 1/16 ≦ Ts ≦ 1, 0 ° ≦ θ (°) ≦ 20 °, 0.25 ≦ R ≦
0.75 where f = −3.26 × 10 ^-1 + 3.19 × 10 ^-2 (1 / °) × θ (°) +1.4
3 × R−4.67 × 10 ⁻⁴ (1 / mm) × We (mm) + 8.39 × 10 ⁻² (1 / mm)
× S (mm) -6.22 × 10 ^-1 × Ts R = ｛(wall thickness at root of pin side lip) / (outer diameter of root at lip)｝ / ｛(wall thickness of pipe body) / (tube body (Outer diameter)｝