Coupled connection for drilling-with-casing operations and tight clearance casing strings in oil and gas wells

Abstract

A tubular system joins tubulars using couplings, which includes continuous internal thread having sections converging and diverging at different angles. A first section's diameter converges linearly inward at a steep angle to a first intersection, a second section converges linearly at a less steep angle from the first intersection to a second intersection, and a third section diverges linearly outward at another angle from the second section. Tangential curvatures at the intersections enhance the coupling's efficiency and fit. Designed for annular clearance inside surrounding casing, the tubulars may have recessed turndowns at the pins. The couplings' outer diameters fit within the annular clearance. Critical cross-sections of the pins and the couplings are balanced to provide tension and pressure ratings of at least 70 percent of the pipe body's ratings.

Description

Coupled Connection for Drilling-With-Casing Operations and Tight Clearance Casing Strings in Oil and Gas Wells-by- Eugene J. MannellaCROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This application is a continuation-in part of U.S. Non-provisional Appl. No. 17 / 840,018 filed June 14, 2022, and claims the benefit of U.S. Provisional Appl. No. 63 / 661 ,766 filed June 19, 2024, both of which are incorporated herein by reference in their entireties.FIELD OF THE DISCLOSURE

[0002] The subject matter of the present disclosure relates generally to threaded connections for joining together tubulars used in oil and gas well exploration and production. More particularly, it relates to couplings for joining individual lengths of casing used in wellbores.BACKGROUND OF THE DISCLOSURE

[0003] Conventional drilling methods to drill an oil and gas well use drill pipe specifically designed for and dedicated to drilling the wellbore. Traditionally, drilling the wellbore involves using a drill bit attached to a drill string to cut through the earth. As the wellbore advances, new sections are drilled with successively smaller drill bits, and the new sections are sealed off with smaller strings of casing until a target depth is achieved. A full casing program consists of multiple telescoping strings of casing, which are cemented in place. After drilling the wellbore, the drill pipe can be transported to another wellsite to drill another well and can be used until it is worn out.

[0004] In contrast to the above procedure, casing can be used for both drilling the well and casing off the open hole. The procedure is commonly referred to as "Drilling With Casing" (DWC) or “Casing while Drilling” (CwD). In this method, the casing itself is used as the drill string and has a drill bit attached to its bottom end. Two types of drill bits are commonly used in drilling with casing operations. One is a retrievable bit, and the other is expendable, is drilled out, and is left behind. Drilling with casing can reduce drilling costs, streamline the drilling process, enhance wellbore stability, and reduce the risks associated with conventional drilling methods.

[0005] In one advantage, for example, drilling with casing can improve wellbore stability. In traditional drilling, the drilled hole remains exposed for a period before thecasing is installed, which can lead to wellbore instability, especially in formations prone to collapse or in high-pressure environments. By using the casing as the drill string, the wellbore is cased immediately as it is drilled, significantly reducing the risk of wellbore collapse. As such, drilling with casing makes it possible to penetrate trouble zones successfully, which may not be possible using “conventional” methods. This immediate casing also minimizes the loss of drilling fluids and reduces the risk of formation damage, which can occur when fluids interact with the formation.

[0006] In another advantage, drilling with casing reduces the time required to drill and complete a well. In conventional drilling, the process of drilling, withdrawing the drill string, and then running the casing into the well can be time-consuming. By combining these steps into a single operation, drilling with casing streamlines the entire process, leading to faster well completion times. This efficiency is particularly beneficial in complex drilling environments, such as in deepwater or unconventional reservoirs, where drilling risks and costs are higher.

[0007] Environmental and safety aspects are also improved with drilling with casing. The reduced exposure time of the wellbore means there is less chance of blowouts, one of the most significant risks in drilling operations. This reduction in risk contributes to a safer working environment for the rig crew. Additionally, since there is less handling of drilling fluids and cuttings, there is a decreased environmental footprint. Finally, drilling with casing can also reduce the number of trips in and out of the well, leading to lower emissions and less wear and tear on drilling equipment.

[0008] In conventional casing usage, the casing and its connections are subjected only to static loads consisting of tension, compression, bending, pressure (internal and external), and any combination thereof. In drilling with casing usage, the casing and connections are not only subject to all of the listed static loads, but they are also subject to cyclic, dynamic loads such as vibration, slip-stick, rotational bending due to rotating the casing and advancement downhole while drilling the wellbore. For example, the casing can be rotated at rotational speeds ranging on the conservative side from about 30 to 120 RPM (revolutions per minute) for drilling operations and 15 to about 40 RPM for advancement to target operations and during cementing operations after the string is fully deployed. Of course, the rotational speed for the casing can vary depending on the implementation, the type of formation being drilled, the casing diameter, the casing material, specific drilling conditions, drilling fluids, bit type, etc. In general, the rotational speeds for the casing would tend to be lower compared to conventional drilling becauserotating the casing at higher speeds can cause wear and fatigue, potentially leading to failure of the casing.

[0009] As the casing rotates and advances down the wellbore the casing string is subject to cyclic fatigue loads. For example, the couplings have a larger outside diameter than the casing and contact the wellbore wall, which causes side impacts and abrasion to the outside diameter surface. With enough wear, the coupling outside diameter will erode reducing the coupling wall thickness. Under certain circumstances, this can lead to a failure in the coupling. Connections deployed for drilling with casing or rotating to achieve target are also known to fail in the pin member a few threads inside the coupling bearing face. This is the area of flexure that experiences the highest stress reversals during rotating operations. Employing the taper change in the prior art reduces bearing stress between the coupling thread crest and the pin tread root. This reduction in bearing stress greatly enhances the fatigue life of the connection. Then, once the casing is set and cemented in the well, dynamic loads cease, and the casing remains subject to all the static loads mentioned above.

[0010] Experience to date with drilling with casing has demonstrated a need for a more robust, yet economical casing connection to withstand the additional rigors of dynamic loading and frictional wear caused by rotating the casing string while drilling or by rotating the string for target achievement. One solution in the prior art is disclosed in U.S. Patent No. 7,347,459 by the same inventor.

[0011] For example, Fig. 1A illustrates a side view, partially in cross-section, of two tubular sections 10a-b joined using a prior art coupling 20 according to the prior art, and Fig. 1B illustrates a side view, partially in cross-section, of two tubular sections 10a-b joined using an alternative coupling 20 of the prior art. Both prior art couplings 20 have internal threads 22a-b that thread to external threads 12a-b on the pins of the tubular sections 10a-b. The prior art coupling 20 in Fig. 1B has an internal, center reinforcing cross-section or ring 27, while the prior art coupling 20 in Fig. 1A does not. The coupling 20 is machined from a single blank, which is cut from a heavy-wall steel tube known in the industry as coupling stock.

[0012] The internal coupling threads 22a-b have multiple tapered sections (S1 , S2, S3). The transition from one taper section (S1 , S2, S3) to another uses a simple intersection of the straight-line tapers. In this prior art coupling 20, the multiple taper sections (S1 , S2, S3) are provided to reduce circumferential hoop stresses and soften the resulting longitudinal hoop stress distribution through thinner cross-sections of boththe coupling 20 and the pin of the pipes 10a-b. This has been achieved by reducing the coupling cross-section relative to the pin cross-section, which reduces localized thread interference and the possibility of thread galling during connection assembly.

[0013] Nevertheless, new drilling technologies and advancements in drilling operations have allowed operators to drill wells with significantly longer lateral sections, which may exceed 22,000 ft (4 miles) in some formations. And every year, operators work to achieve even longer lateral sections. With every increase in lateral length beyond about 12,000 to 15,000 ft, the challenges of successful casing deployment increase exponentially. While drilling with casing is still a viable process, it is less common compared to the process of rotating casing to achieve a target depth in long lateral wells. In this process, initial sections of the well are drilled using a conventional drill string, which can be rotated and steered to follow the planned path. This includes drilling down to the kickoff point where the well begins to deviate from vertical.

[0014] After reaching a predetermined depth, developing the curved and lateral sections of the wellbore, casing is then run into the well. Casing running tools (CRT) used at the rig can both rotate and reciprocate the casing as it is run into the hole. The rotation reduces friction, which is particularly effective in long lateral wells where reaching extended depths are hindered by increased friction and complexity of advancing along the irregular wellbore.

[0015] For example, many of today’s wells are directional with a vertical section, a curved section, and then a long horizontal lateral. Lateral sections in contemporary wells are typically 5,000 to 12,000 ft. Although, so-called, super-lateral wells have been drilled with 22,000 ft (4 mile) horizontal sections. Once the drill string is pulled from the open hole, casing is deployed by making up individual lengths of casing and lowering them into the open hole. Getting the casing down into the well, around the curved section, and out to the end of a long lateral section is often difficult. To assist casing deployment, operators may have to reciprocate (push and pull) and / or rotate the casing to assist advancement to the end of the long lateral section. Working casing strings in this manner to achieve target depths or to free stuck pipe imparts significant variable load combinations that are cyclic in nature that can potentially cause early and unexpected failure of connections by fatigue, helical buckling, and other failure mechanisms.

[0016] In the oil and gas wells, strings of casing tubulars can telescope one inside another as they are deployed into the well. A typical casing program starts with largediameter tubulars are followed by successively smaller diameter tubulars, the individual casing strings within the well can include conductor pipe, surface casing, intermediate casing, production casing, liners, and production tubing used to case off the wellbore and provide a conduit for producing oil and gas.

[0017] Certain regulatory agencies require a minimum annular clearance between successive casing strings deployed in wells drilled within their jurisdictions. To meet these minimum annular clearances, operators typically use flush or semi-flush integral joint “premium” connections to fit in the tight clearances between the successive casing strings in a well.

[0018] For example, Fig. 1C illustrates a typical flush joint connection assembly 11A available in the art. The pipe bodies of casing tubulars 10a-b have outside diameters OD and have bores 15a-b with inner diameters ID. A male pin end 14a of one of the tubulars 10a is shown threaded to a female box end 14b of the adjoining casing tubular 10b. Both tubulars 10a-b in a casing string would have male and female ends joined together to create a casing string during installation in a well. The male pin end 14a is threaded along a reducing diameter from the outer diameter OD to an intermediate diameter D. Likewise, the female box end 14b is threaded along an increasing diameter from the intermediate diameter D to the outer diameter OD. At the end of the male pin end 14a is a short reduced and finely machined OD cylindrical surface 18a which is designed to radially interfere with a complimentary finely machined ID cylindrical surface 18b of the mating female box end 14b. The surface finishes and mating contact pressure of these surfaces 18a and 18b create a metal-to-metal seal when the connection assembly 11A is properly assembled to effectively contain high-pressure gas. Connections with metal-to-metal seals are often referred to as “premium” connections in the industry.

[0019] As can be seen, the mating threads for the male and female ends 14a-b are machined into the bodies of the tubulars 10a-b so the male and female threaded ends 14a-b each share the casing wall thickness. For this reason, the flush joint connection assembly 11 A generally has low tension efficiency relative to the pipe bodies of the adjoined tubulars 10a-b. Additionally, a distal face 16a of the male pin end 14a, often referred to as the pin nose, may or may not engage with an inner face 16b of the female box end 14b. Without engagement, the connection assembly 11A has no additional capacity for torque resistance greater than that needed to create radial interferences between the male pin end 14a and female box end 14b and to energize the metal-to-metal seal (if present) during assembly of the connection assembly 11 A. With engagement, torque resistance increases slightly. However, the face 16a and mating surfaces result in relatively small cross-sectional nose bearing areas because they each consume a portion of the shared casing wall thickness and therefore provide minimal additional torque resistance.

[0020] In another example, Fig. 1 D illustrates a typical semi-flush joint connection assembly 11 B available in the art. For the semi-flush connection assembly 11 B, one casing tubular 10a has a reduced pin end 14a, and an adjoining casing tubular 10b has an expanded or upset box end 14b, which threads to the reduced pin end 14a. Both tubulars 10a-b in a casing string would have pin and box ends joined together to create a string during installation in a well.

[0021] The pipe bodies of both casing tubulars 10a-b generally have an outer diameter OD and have a bore 15a-b with an inner diameter ID. The reduced pin end 14a is threaded along a reducing diameter from the outer diameter OD to an intermediate diameter D2. A stepped shoulder 17a may or may not be provided at an intermediate transition. The expanded box end 14b of the other tubular 10b has an expanded or upset outer diameter D3 that is greater than the outer diameter OD of the pipe body. The expanded box end 14b may or may not include a complimentary intermediate stepped shoulder depending on the pin member configuration. The expanded box end 14b with no intermediate stepped shoulder is threaded along a constantly decreasing diameter transitioning from Di to D2 to engage the reduced pin end 14a threaded along a complimentary decreasing diameter from the outer diameter OD to a diameter D2. When a stepped shoulder is provided in an intermediate transition in the expanded box end 14b, the box end 14b is threaded along the increasing diameters D2 to Di to engage the stepped shoulder 17a of the pin end 14a threaded along complimentary decreasing diameters Di to D2.

[0022] Expanded box end 14b and reduced pin end 14a share a portion of the casing wall thickness. For this reason, the semi-flush joint connection assembly 11 B generally has low tension efficiency relative to the pipe bodies of the adjoined tubulars 10a-b. The distal face 16a of the pin end 14a may or may not shoulder inside the box end 14b, and the distal face 16b of the box end 14b fits outside the other casing tubular 10a. At the end of the male pin end 14a is a short reduced and finely machined OD cylindrical surface 18a which is designed to radially interfere with a complimentary finely machined ID cylindrical surface 18b of the mating female box end 14b. The surfacefinishes and the mating contact pressure of these surfaces 18a-b create a metal-to- metal seal when the connection is properly assembled to seal high-pressure gas. Connections with metal-to-metal seals are often referred to as premium connections in the industry.

[0023] Like the flush joint, the ends 14a-b of the semi-flush joint 11 B each consume a portion of the casing wall thickness. However, more wall of the pipe body is available for each tubular 10a-b due to the expanded box end 14b and reduced ends 14a. Still, the semi-flush connection 11 B has reduced tension efficiencies relative to the pipe body but typically provide greater performance than a flush joint connection assembly 11 A. Similarly, if distal face 16a of the pin end 14a engages with distal face 16b of the box end 14b, there is an increase in torque capacity. If there is an intermediate step shoulder 17a, engagement torque capacity is also increased.

[0024] As is expected, these flush and semi-flush joint connections 11 A-B are expensive due to manufacturing costs and limited industry availability (few designs and suppliers). In many instances, the intended service for these flush and semi-flush joint connections 11A-B does not comport well with performance needs and the price required for a given installation.

[0025] Although existing casing connections may be useful and beneficial, ever- increasing demands are being placed on casing connections used in drilling with casing operations and / or rotating casing to aid target achievement to meet the new drilling technology demands, and difficulties with casing advancement in drilling operations. Moreover, standard American Petroleum Institute (API) threaded connections have a very loose tolerance requirement. For example, the allowable make-up variation is 0.575 inches (i.e., 2.88 turns) between the minimum and maximum make-up positions. For these reasons, the subject matter of the present disclosure is directed to addressing issues associated with drilling with casing, rotating to achieve the target landing location in the well and particularly to addressing the connections used to join each length of casing together.

[0026] The subject matter of the present disclosure is directed to overcoming, or at least reducing the effects of, one or more of the problems set forth above.SUMMARY OF THE DISCLOSURE

[0027] In one configuration, a coupling is used for joining tubulars. The coupling comprises a body having a first end and a second end and defining a boretherethrough. At least a first portion of the bore defines a first continuous internal thread, which has a first section, a second section, and a third section. The first section is disposed toward the first end, the second section is connected with the first section at a first intersection, and the third section connected with the second section at a second intersection.

[0028] An internal diameter of the first continuous internal thread for the first section converges linearly inward toward the bore at a first angle from a first point to the first intersection. The internal diameter for the second section converges linearly inward toward the bore at a second angle from the first intersection to the second intersection. The second angle is less than the first angle. Finally, the internal diameter for the third section diverges linearly outward from the bore at a third angle from the second intersection to a second point. The internal diameter of the first continuous internal thread defines at least one of: (i) a first curvature transitioning between the first and second angles at the first intersection of the first and second sections, and (ii) a second curvature transitioning between the second and third angles at the second intersection of the second and third sections. In one arrangement, the first curvature is tangential to the first and second angles, and / or the second curvature is tangential to the second and third angles.

[0029] In another configuration of the present disclosure, a tubular system comprises a plurality of tubulars and a plurality of couplings. Each of the tubulars has a pipe body with a first outer diameter and defines a first bore with an inner diameter. Each of the tubulars has pins disposed on ends of the tubular, and the pins have external thread. The couplings are configured to join the tubulars together and are configured as described above.

[0030] In one arrangement, an annular clearance limit for the tubular system downhole constrains how large the first outer diameter of the tubulars can be and constrains how large a given second outer diameter of the couplings can be. The first outer diameter of the tubulars is essentially defined by the implementation, and the given second outer diameter of the couplings is limited by how much annular clearance by regulation must be maintained inside a surrounding casing ID or wellbore ID. Each of the pins of the tubulars has a turndown recessed from the first outer diameter to a first intermediate outer diameter smaller than the tubulars’ first outer diameter. Each of the pins then tapers inward from that first intermediate outer diameter at the turndown to a second intermediate outer diameter at a nose of the pin. Because the couplings aregiven the second outer diameter (OD) to meet the annular clearance limit set by regulation, the turndown allows the inside diameter at the box ends of the couplings to be reduced, which in turn increases the cross-section of the box ends of the couplings. In effect, a portion of each pin on the tubulars is shared with the box ends on the couplings to provide a stronger coupling. The pin and coupling cross-sections are therefore balanced to produce a matched strength design to meet target connection and performance properties.

[0031] The tubulars define first critical cross-sections between the first outer diameter of the pipe body and the inner diameter of the pipe body, between the first intermediate outer diameter of the turndown and the inner diameter of the pipe body, and the second intermediate outer diameter of the pin and the inner diameter of the pipe body. Meanwhile, the couplings define second critical cross-sections between the second outer diameter and the internal diameter of the coupling. The first and second critical cross-sections can be configured to maximize at least one of: (i) the connection tension resistance of the connection relative to the pipe tension resistance of the pipe body, and (ii) the connection pressure resistance of the connection relative to the pipe pressure resistance of the pipe body. For example, the connection tension resistance can be at least 70 percent of the pipe tension resistance, and the connection pressure resistance can be at least 70 percent of the pipe pressure resistance.

[0032] The foregoing summary is not intended to summarize each potential embodiment or every aspect of the present disclosure.BRIEF DESCRIPTION OF THE DRAWINGS

[0033] Fig. 1 A illustrates a side view, partially in cross-section, of two pipes joined using a coupling according to the prior art.

[0034] Fig. 1B illustrates a side view, partially in cross-section, of two pipes joined using an alternative coupling of the prior art.

[0035] Fig. 1C illustrates a typical flush joint connection available in the art.

[0036] Fig. 1D illustrates a typical semi-flush joint connection available in the art.

[0037] Fig. 2A illustrates a cross-sectional view of a connection assembly according to the present disclosure joining two tubulars together.

[0038] Fig. 2B illustrates a cross-sectional view of another connection assembly according to the present disclosure joining two tubulars together.

[0039] Fig. 3A illustrates a side view, in partial cross-section, of the connection assembly of Fig. 2A, further including an unthreaded extension.

[0040] Fig. 3B illustrates a side view, in partial cross-section, of the connection assembly of Fig. 2B, further including an unthreaded extension.

[0041] Fig. 4A illustrates a schematic of continuous internal thread (not to scale) having three sections with adjacent sections intersecting at a point according to the present disclosure.

[0042] Fig. 4B illustrates another schematic of the continuous internal thread (not to scale) having three sections where each section transitions to the adjacent section through a tangential curve according to the present disclosure.

[0043] Figs. 5A-5B schematically illustrate generic first and second types of tooling for making a thread form.

[0044] Fig. 6A schematically illustrates a representation of a helical profile for two tapered sections meeting at an intersection according to the prior art.

[0045] Fig. 6B schematically illustrates a cross-section of the representation in Fig. 6A.

[0046] Fig. 7A schematically illustrates a representation of a helical profile for two tapered sections meeting at a tangentially curved transition according to the present disclosure.

[0047] Fig. 7B schematically illustrates a cross-section of the representation in Fig. 7A.

[0048] Fig. 8A illustrates a schematic analysis of force during mill makeup of the prior art coupling to the “mill end” of a tubular.

[0049] Fig. 8B illustrates a schematic analysis of force during field makeup of the “field end” of another tubular to the prior art coupling.

[0050] Fig. 9A illustrates a schematic analysis of force caused by mill makeup of the disclosed coupling to the “mill end” of a tubular.

[0051] Fig. 9B illustrates a schematic analysis of force caused field makeup of the “field end” of another tubular to the disclosed coupling.

[0052] Fig. 10A illustrates a cross-sectional view of another connection assembly according to the present disclosure, joining two tubulars together with a coupling.

[0053] Fig. 10B illustrates a cross-sectional view of yet another connection assembly according to the present disclosure, joining two tubulars together with a coupling.

[0054] Fig. 11 illustrates a cross-sectional view of a connection assembly according to the present disclosure for joining two tubulars together with a coupling in a slim hole coupled connection.

[0055] Fig. 12 illustrates a schematic view of the coupling of the disclosed connection assembly for the slim hole coupled connection having the multiple taper sections for the thread.

[0056] Fig. 13 illustrates a schematic view of a reduced pin of the disclosed connection assembly for the slim hole coupled connection.

[0057] Fig. 14 illustrates a schematic view of the coupling of the disclosed connection assembly for the slim hole coupled connection.

[0058] Figs. 15A-15B illustrate example configurations for achieving a connection assembly for a slim hole coupled connection of the present disclosure.

[0059] Fig. 16 illustrates the connection assembly for the slim hole coupled connection schematically showing stresses and displacements due to the use of high radial interference mating threads.DETAILED DESCRIPTION OF THE DISCLOSURE

[0060] Fig. 2A illustrates a cross-sectional view of a connection assembly 50 according to the present disclosure. The connection assembly 50 is used in a system to join tubulars 60a-b together to make a casing string for downhole deployment.(References made herein to tubular, casing, or pipe apply equally to one another.) As shown, a portion of a casing or tubing string having two tubular sections 60a-b are interconnected with a coupling 70. The tubular sections 60a-b can be casing sections, pipe, tubing, or other tubular components. The coupling 70 can be a body or member that is a hollow cylindrical. The casing sections 60a-b have pins 64a-b defining external thread 65a-b, which mate with internal thread 75a-b of the coupling 70. The threaded pins 64a-b can contain standard American Petroleum Institute (API) Buttress Threads 65a-b with a constant taper. These threaded pins can also contain alternate API Threads, such as 8-round, or other industry-standard or proprietary thread forms. The faces or noses 66 of the two pins 64a-b can have square cut ends to furnish maximum bearing face when butted together at a center of the coupling 70.

[0061] Under standard practice, threads of a standard coupling would have an identical thread taper with the mating thread form (API Buttress Thread Form in this discussion) on the pin threads 65a-b so uniform radial thread interference can beproduced through the full length of the thread profile. When the connection is assembled, this radial thread interference creates the contact pressure along the mating threadform interface that provides the desired sealing capabilities in the assembled connection.

[0062] As can be seen, the threads 65a-b on both the pins 64a-b and the threads 75a-b on the coupling 70 both taper, which results in variable cross-sections along the thread profile of each member. A thinner cross-section occurs at the faces or noses 66 of the pins 64a-b, associated with similar thinning cross-sections at the coupling’s ends or bearing faces 71a-b. When the connection is assembled, the thinner cross-sections of the respective pin faces or noses 66 and coupling’s ends 71a-b are opposite the heavier cross-sections of the mating member. The relative pin and coupling crosssections at these locations therefore are imbalanced at the thinner ends of both members.

[0063] As expected, a uniform taper between the external threads 65a-b on the pins 64a-b and internal threads on a standard coupling would produce uniform interference along the thread profile. However, as pointed out above, the cross-sections of the mating members vary along the thread profile. Therefore, if the interference between the threads is uniform, but the cross-sections behind the threads are variable, then the resulting hoop stresses created in the cross-sections must also be variable, graduating from low stresses in the thicker part of the cross-sections in each member to high stresses in the thinner part. Indeed, Finite Element Analysis (FEA) shows that, after assembly, the hoop stresses in the thinner cross-sections of both the pins 64a-b and a standard coupling can approach and, with certain wall thicknesses of pins 64a-b, exceed the yield strength of the material used (e.g., steel). In addition to the negative impact of exceeding the yield strength in the thinner portions of the mating members, this differential yielding at the thin vs. thick cross-sections also causes differential movement between the threads at these same high stress points. This differential movement, at the high stress points, in turn results in thread galling in both the standard coupling threads 75a-b and the pins 64a-b particularly at the taper transition points T1 and T2. It is also anticipated that these same high stressed areas, particularly at the run-out threads at P2 of the pins 64a-b, can result in fatigue failure when the standard couplings are used in the drilling with casing or with rotating operations employed for target achievement when deploying casing in long lateral well sections.

[0064] In the cross-section shown, a shortened API Buttress Threaded coupling 70 connects two Buttress Threaded pins 64a-b that abut at the center of the coupling 70. This shortened and multiple tapered coupling 70 is designed to: (1 ) moderate concentrated mating thread interface bearing loads that cause high stresses previously outlined, (ii) reduce thread galling in areas of high stress, (ii) maintain compatibility with standard API Buttress threaded pins 64a-b, and (4) create a high torque connection by butting the pins faces or noses 66 at the center of the coupling 70.

[0065] To accomplish these objectives, the internal thread of the coupling 70 has varied or modified tapering at the areas of high stress (i.e., the areas of cross-sectional imbalances at coupling ends 71a-b and at pin faces or noses 66). In particular, the internal thread 75a-b in coupling 70 is segmented into sections (S1 , S2, S3). When the connection is assembled, the imbalance between the noted cross-sections and any resulting excessive hoop stresses in the thinner cross-sections can be addressed using the varying taper sections (S1 , S2, S3) for the internal threads 75a-b of the coupling 70.

[0066] Accordingly, the connection assembly 50 for joining the casing sections 60a-b includes the coupling 70, which can be a cylindrical body or member, having a first (field) end 71a and a second (mill) end 71 b and defining a bore 72 therethrough. At least a first portion of the bore 72 defines a first continuous internal thread 75a, which has a first section (S1 ), a first transition (T1 ), a second section (S2), a second transition (T2), and a third section (S3). The first section (S1) is disposed at the first (field) end 71a, the second section (S2) is connected with the first section (S1 ) by the first transition (T1 ), and the third section (S3) is connected with the second section (S2) by the second transition (T2). Furthermore, a second portion of the bore (72) can also define a second continuous internal thread 75b mirroring the first continuous internal thread 75a.

[0067] The internal diameter 76a-b of the continuous internal threads 75a-b varies along the axial length of the coupling 70 so the sections (S1 , S2, S3) have different tapers for the internal threads 75a-b from one section to the next. In particular, the taper in the second section (S2) can be maintained at the API standard taper to directly match the single taper machined on pin threads 64a-b. The matching pin and coupling tapers in S2 take advantage of the relative balance of the coupling and pin crosssections. However, the taper in the first section (S1 ) can be greater than the taper in the second section (S2), and the taper in the second section (S2) can be greater than the taper in third section (S3). The varied tapers of the internal thread 75a-b in thecoupling 70 relative to the uniform tapers of the external thread 65a-b on the pins 64a-b directly reduce the bearing loads imparted by the coupling thread crests on the pin thread roots in the mating thread elements in areas of the connection with unbalanced cross-sections in sections S1 and S3 (i.e., starting at points T1 and T2 at the transitions and varying along each section (S1 and S3) on the pins to the ends at P1 and to the coupling’s ends 71a-b at P2). Employing multiple taper sections (S1 , S2, S3) reduces the contact pressure in high areas of localized stress for sections (S1 ) and (S3) and thus mitigates problems of high stress, thread galling, and fatigue failure.

[0068] In one configuration, the internal thread 75a-b on each end of a 7-inch API Buttress Threaded coupling 70 is divided into three sections (S1 , S2, S3) as previously described. The lengths and tapers for each section (S1 , S2, S3) can be given for one particular casing OD and pipe wall thickness combination are as follows:Section Length (in) Taper (in / in)51 1.784 0.0752552 1.716 0.0625053 1.125 0.05556

[0069] The coupling 70 can be shortened (e.g., by approximately 3 / 4 inch in some configurations) by removing what is commonly known as the "J" area between the two pins 64a-b in standard API Buttress Connections manufactured in accordance with API Specifications as enumerated in API Specifications 5CT and 5B. Eliminating the "J" area allows the faces or nose 66 of the two pins 64a-b to butt one another at the coupling’s center when assembled to the power tight position, thereby providing high torque resistance needed for drilling with casing or casing rotating operations employed during casing deployment.

[0070] Fig. 2B illustrates a cross-sectional view of another connection assembly 50 according to the present disclosure joining two tubulars (casing sections) 60a-b together. The coupling 70 has an internal reinforcing cross-section or ring 80 at the center or opposite the "J" area. In this connection assembly 50, the face or nose 66 of each pin 64a-b abuts an internal square shoulder 86 at the heavy cross-section or ring 80, thereby providing high torque resistance needed for drilling with casing or casing rotating operation employed during casing deployment. The coupling 70 can be machined from a single piece of steel (i.e., coupling stock).

[0071] Fig. 3A illustrates a side view, in partial cross-section, of the connection assembly 50 of Fig. 2A, further including an optional unthreaded extension 90. The coupling 70 can include the optional unthreaded extension 90 integrally machined onone (mill) end 71 b of the coupling 70. The extension 90 can provide a sacrificial wear sleeve to protect the main body of the coupling 70 as the casing is rotated down the wellbore in an abrasive environment. The wear extension 90 would have the same outside diameter as the coupling 70 with the inside diameter being slightly larger than the casing sections 60b so as to slip over the casing section 60b when the connection is assembled. As an option, the wear extension 90 can be hard banded if excessive abrasion is anticipated.

[0072] The inside diameter 92 of the wear extension 90 can be uniform from the face91 for a specific distance toward the center of the coupling 70. The inside diameter 92 can then be tapered outward relative to the outside diameter of the coupling 70. This permits a threading tool to cut perfect (full formed) thread (e.g., 75b) over the entire coupling thread length without cutting into the inside diameter of the sacrificial wear extension 90. Elimination of machine marks in the inside diameter of the wear extension 90 near the coupling’s internal threads 75b reduces the possibility of fatigue failures in the sacrificial wear extension 90.

[0073] Fig. 3B illustrates a side view in partial cross-section view of the coupling 70 of Fig. 2B, also including an optional unthreaded extension 90. Again, the optional unthreaded extension 90 can be integrally machined on one (mill) end 71b of the coupling 70. The extension 90 provides a sacrificial wear sleeve to protect the main body of the coupling 70 as the casing is rotated down the wellbore. The wear extension 90 can have the same outside diameter 94 as the coupling 70, and the inside diameter92 can be slightly larger than the casing section 60b so as to slip over the casing section 60b when the connection is assembled. As an option, the wear extension 90 can be hard banded on part of its external surface should excessive abrasion be anticipated.

[0074] As before, the inside diameter 92 of the wear extension 90 can be uniform from the face 91 for a specific distance toward the center of the coupling 70, then can tapered outward relative to the outside diameter of the coupling 70. This permits a threading tool to cut perfect (full formed) thread (e.g., 75b) over the entire coupling thread length without cutting into the inside dimension of the sacrificial wear extension 90. Elimination of machine marks in the wear extension 90 near the coupling threads reduces the possibility of fatigue failures in the sacrificial wear extension 90.

[0075] Having an understanding of the connection assembly 50 of the present disclosure, discussion turns to additional details about the continuous internal thread ofthe coupling 70. Fig. 4A illustrates a schematic of continuous internal thread 75 having three sections (S1 , S2, S3) according to the present disclosure. Fig. 4B illustrates another schematic of the continuous internal thread 75 having curvatures at the transitions (T1 , T2) between the sections (S1 , S2, S3) according to the present disclosure.

[0076] In these schematics, an internal diameter 76 mid-way between the thread roots and crests of the continuous internal thread 75 is shown varying along the axial length so the sections (S1 , S2, S3) have different tapers for the internal thread 75 from one section to the next. For simplicity, the profile of the threading for the continuous internal thread 75 is not shown. For comparative purposes, the profile and taper of external thread 65 for a pin of a casing section is illustrated to the side of the contour of the continuous internal thread 75. (It should be noted that the tapers in Figures 4A and 4B are greatly exaggerated relative to the external thread tapers of pin thread 65 for illustrative purposes. Additionally, it will be understood that the 2-Dimensional depictions shown here in Figs. 4A-4B as well as in any other drawings are simply a convenient way to depict the complicated 3-Dimensional geometry of the present disclosure. The transition curves between sections of the thread are actually machined on a helix, which is only partially depicted in Figs. 7A-7B.)

[0077] In the contour of the continuous internal thread 75, the internal diameter 76 for the first section (S1 ) converges linearly inward (i.e., toward the axial centerline A in the coupling’s bore) at a first angle or taper (a) from a first point (p1) to the first transition (T 1 ). The internal diameter 76 for the second section (S2) converges linearly inward at a second angle or taper (P) from the first transition (T1 ) to the second transition (T2). This second angle (P) is less than the first angle (a). The internal diameter 76 for the third section (S3) diverges linearly outward (i.e., away from the axial centerline A in the coupling’s bore) at a third angle or taper (- / ) from the second transition (T2) to a second point (p2). (Because the third angle is diverging, it is labelled as a negative angle (- / ) in the discussion).

[0078] In contrast to the continuous internal thread 75, the external threads 65 for the pin of the casing section may be made to standard API Buttress specifications without modifications to length or taper. This external thread 65 is schematically illustrated to the side of the continuous internal thread 75. As noted, the continuous internal thread 75 is configured to thread to a pin having a single taper angle (P’).Accordingly, at least one of the first angle (a), the second angle (P), and the third angle (- / ) of the continuous internal thread 75 is approximate to the single taper angle (P’) of the pin thread 65. In particular, the second angle (P) of the second section (S2) can be approximate (i.e., should be matched within acceptable machining tolerances) to the single taper angle (P’) of the external thread 65 on the pin (64). In addition, the third angle (- / ) diverging outward can reduce galling and can reduce the chances of radial buckling in some instances when the pin (64) is threaded in the coupling (70) because the coupling (70) at the third section (S3) has its thickest wall whereas the pin (64) would be at its thinnest. Moreover, the third angle (- / ) diverges outward to reduce the thread interference, which can actually bring the thread interference comparable to that in standard API connections using the same thread form.

[0079] The taper angles (a, p, - / ) defined for the sections (S1 , S2, S3) can be linear because the more linear sections (S1 , S2, S3) can maximize the lengths (L1 , L2, L3) of the sections (S1 , S2, S3), and especially the second section (S2) that has the same taper on both members, thereby maximizing sealing integrity. The section lengths (L1 , L2, L3) and the taper angles (a, p, - / ) employed can be designed to reduce contact pressure in areas of the connection where cross-sections are imbalanced. As will be appreciated, the actual lengths (L1 , L2, L3) and taper angles (a, p, - / ) for the sections (S1 , S2, S3) can depend on the constraints (relative coupling and pin wall thicknesses and diameters) of a particular implementation.

[0080] Instead of an angular change (angle) at the intersection (11 , I2) between the tapers of the sections (S1 , S2, S3), the first and second transitions (T1 , T2) each define a curvature that transitions from one taper section to the next. (Although both transitions (T1 , T2) have curvatures, only one of them may have a curvature in another configuration).

[0081] A shown in detail in Fig. 4B, for example, the internal diameter 76 of the continuous internal thread 75a for the first transition (T1 ) defines a first curvature transitioning from the first angle (a) to the second angle (P), and the internal diameter 76 of the continuous internal thread 75 for the second transition (T2) defines a second curvature transitioning from the second angle (P) to the third angle (- / ).

[0082] Each of the transitions (T1 , T2) can be a curved, tangential transition. For example, the first angle (a) and the second angle (P) are tangential to the first curvature of the first transition (T 1 ); and the second angle (P) and the third angle (- / ) aretangential to the second curvature of the second transition (T2). These curved transitions (T1 , T2) can further minimize the possibility of thread galling and yield significantly increased connection fatigue resistance by reducing the normal force (or bearing load) imparted between the coupling thread crests and pin thread roots at the transitions T1 and T2 during assembly of the connection improving on the features in the prior art. Moreover, the arrangement having the curved transitions (T1 , T2) effectively converts a point load to a load distributed over a defined length.

[0083] Manufacture of the continuous internal thread 74 having the taper sections (S1 , S2, S3) and the curved transitions (T1 , T2) can be implemented using a threading program of a Computer Numerical Control (CNC) machine used for manufacturing couplings and threaded connections.

[0084] In general, forming the taper of threads is set by the type of threading tool used. Two types of tooling for threading tools are commonly available today. Fig. 5A schematically illustrates a first type of tooling 40a for making a thread form. This first type of tooling 40a is used for machining internal thread 35 on a component 30. The threading or cutting tool 42 has a single thread form on each corner that can cut the internal thread 35 of the component 30 along a programmed path that axially advances the tool 42 against the rotating component 30 along the thread length while changes to the radius (internal diameter) are made that removes material to form the finished threads. The final, machined threads are formed after multiple threading passes of the tool 42 during manufacturing. As shown, the cutting tool 42 can be triangular with three corners. As the cutting tool 42 wears, it is turned to use another corner until all comers are worn out.

[0085] Fig. 5B schematically illustrates a second type of tooling 40b, which is used for making an external thread 37 on a component 30. In this second type of tooling 40b, the threading or cutting tool 44 has three or more, co-linear thread forms causing varying depths of cuts to form the desired thread form for the external thread 37. This second type of tooling 40b requires fewer passes to cut the external thread 37 along the specified length.

[0086] In the prior art, such as in US 7,347,459, the changes in internal thread taper were achieved with the first type of tooling by making small incremental changes in radius as the threading tool axially advances along the length. The incremental changes create a small step at the intersections (11 , I2) in individual finished threads in the variable taper sections (S1 , S2, S3). These steps at the intersections (11 , I2) can beseen using an optical comparator at a 50X magnification. The prior art techniques have attempted to manage this process so these inherent steps in the thread caused by fixed tooling characteristics do not adversely impact the fit, form, or function of the mating threads.

[0087] In contrast to the prior art, the techniques of the present disclosure “soften” the inherent small steps from one tapered segment to another in the finished threads 75 by using the tangentially, curved transitions (T1 , T2). The techniques of the present disclosure seek to improve the prior art by further reducing normal forces in areas prone to thread galling and fatigue failure initiated by deployment operations involving drilling with casing and / or reciprocating and rotating the casing either singularly or in combination to aid landing-target achievement for the casing in horizontal laterals.

[0088] Further details related to the prior art intersections and the curved transitions of the present disclosure are discussed with reference to Figs. 6A to Fig. 7B.

[0089] In particular, Fig. 6A schematically illustrates a representation of a helical profile for two tapered sections meeting at an intersection according to the prior art, and Fig. 6B schematically illustrates a cross-section of the representation of the helical profile in Fig. 6A. The threads of the thread form that would be defined along this helical profile are not shown, and the tapered sections that meet at the intersection of adjacent tapers would extend additional lengths in the axial direction A.

[0090] For comparison, Fig. 7A schematically illustrates a representation of a helical profile for two tapered sections meeting at a curved transition according to the present disclosure, and Fig. 7B schematically illustrates a cross-section of the representation of the helical profile in Fig. 7A. As will be appreciated, the features of a curved transition between tapered sections depicted in Figs. 4A-4B have been illustrated with two- dimensional representations. When the curved transition between tapered sections is formed on a coupling as disclosed herein, however, it will be appreciated that the curved transitions between the tapered sections are made on a three-dimensional mechanical component. The teachings of the present disclosure overcome the complexity of this challenging process to produce smooth transitions from one threaded section to another within the helical thread.

[0091] Again, the threads of the thread form that would be defined along this helical profile are not shown, and the first taper section enters a curved transition tangential to the first taper which exits the curved transition that is tangential to the second taper.Whether entering or exiting the tangentially curved transition the respective tapers extend additional lengths in opposite axial directions A.

[0092] These schematic representations of the helical profiles in Figs. 6A-6B and 7A-7B are enlarged to illustrate the difference between the prior and the disclosed configuration. While not presented to scale, the schematic representations are sized consistently for ready comparison. As noted, helical threads (not shown) would follow the helical profile. The change from one tapered section to another tapered section occurs over a small portion of the circumference of the interior threaded surface. In particular, the prior art transition as in Figs. 6A-6B is formed over a much smaller portion of the interior threaded helical surface of the helical profile and may occur over about a quarter ( ) rotation of the threading tool. By contrast, the smooth curved transition of the present disclosure as in Figs. 7A-7B spans a larger portion of the interior threaded helical surface of the helical profile and may have a span (s) over a half (>2) turn to one (1 ) full rotation of the coupling 70 as the coupling 70 is threaded by the cutting tool used to form the features. In general, the span(s) of the helical profile is defined by (i) a relative thickness of the mating components (the coupling 70 and the tubular 60a-b) and (ii) the radius (r) of the curved transition.

[0093] The span of the disclosed curved transition over portion of the interior threaded surface’s circumference can be configured for a given implementation. In general, the curved transition may have a greater span over more of the internal threads when implemented with oversized couplings and / or casing with heavy wall thickness. The delineation between what constitutes smaller and larger sizes in this context may occur at a pipe size with a wall thickness greater than about 0.550 inches.

[0094] As noted, the continuous internal thread 75 has curved transitions (T1 , T2) that are tangent to the merging taper sections (S1 , S2, S3). Considering that helical threads for the continuous internal thread 75 are machined into the coupling body that is a hollow cylinder, each of the curved, tangential transitions (T1 , T2) presents a complicated 3-dimensional challenge for programming and machining. The lengths and radii (R1 , R2) of the curved transitions (T1 , T2) are preferably engineered to improve performance over existing thread arrangements known in the art. In particular, the axial length of each transition (T1, T2) is preferably configured to be within ± 1 to 2 thread pitches on either side of the hypothetical intersections or points (11, 12) between the meeting tapers (a, 0, - / ). As defined, the thread pitch for the continuous internal thread 75 is the measurement of how far the continuous internal thread 75 advances axially inone complete turn about the cylindrical bore of the coupling so the thread pitch can be defined as the reciprocal of the number of threads per unit length. In one configuration, the axial length (a that the helical profile extends can be between about 1 / 4 to 1 / 2 of the thread pitch to about 1 / 2 to 1 of the thread pitch of the continuous internal thread 75. Again, the axial length (a of the helical profile is defined by (i) a relative thickness of the mating components (the coupling 70 and the tubular 60a-b) and (ii) the radius (r) of the curved transition. Moreover, the radius (R1, R2) for each transition (T1 , T2) is also configured so proper thread interference can be maintained for connection axial load and leak resistance.

[0095] In that sense, the continuous internal thread 75 can reduce local bearing loads between the coupling thread crests on the pin thread roots when the connection is fully assembled for improved galling resistance and significantly greater resistance to long-term, low-level cyclic loading. These improvements can give the coupling (70) improved fatigue life during service and can allow for substantially more rotating hours in difficult well conditions. The enhanced utility gives operators a considerably better chance of achieving target depths before fatigue failure when drilling with casing or rotating casing for target achievement in extended reach, horizontal laterals.

[0096] The coupling 70 is a complicated 3-dimensional component machined to have a continuous (helical) internal thread 75. In a 2-dimensional situation as depicted in Fig. 4A, the bearing force developed at the intersections (11 , I2) between sections (S1 , S2, S3) occurs at a single point, which has no area, if the curved transitions (T1 , T2) as disclosed herein are not present. This condition theoretically causes infinite stress due to the load at the single bearing points of the intersections (11 , I2). In the complicated 3-dimensional coupling 70 with a continuous (helical) internal thread 74, the area over the intersections (11 , I2) is quite small but does not occur at a single point as depicted in the 2-dimensional example. However, the actual bearing area over which assembly bearing forces attenuate is relatively small, and the localized bearing force of the coupling thread crest on the mating pin thread root causes localized high stress that contributes to thread galling and reduced fatigue life compared with the present invention.

[0097] The continuous internal thread 75 disclosed in Fig. 4B having the curved transitions (T1 , T2) greatly reduces localized bearing stresses by spreading or attenuating localized bearing forces over a designed axial length, which is curved. Thisprovides an improvement over the prior art and provides operators with increased rotating time before connection endurance limits may be exceeded.

[0098] The continuous internal thread 75 with the curved transitions (T1 , T2) between taper sections (S1 , S2, S3) as in Fig. 4B is expected to have little to no effect on costs while improving downhole utility through increased endurance limit of the connection in service. In practice, the coupling 70 manufactured with this continuous internal thread 75 is expected to exhibit no overt changes during assembly or use after downhole deployment. However, performance improvements relative to fatigue resistance and reduced galling can be significant.

[0099] As noted in the background of the present disclosure, for instance, new drilling technologies and advancements in drilling operations have allowed operators to drill wells with significantly longer lateral sections, which may exceed in some cases 22,000 ft (4 miles). The connection assembly (50) disclosed herein supports drilling these longer lateral sections. The connection assembly (50) disclosed herein can significantly improve the likelihood of successfully deploying casing to the intended target or surviving operations to free stuck pipe in difficult wellbore situations. This feature gives the operator greater utility to work the casing to achieve target depth and / or to free stuck pipe.

[0100] Analysis shows that the curved, tangential transitions (T 1 , T2) reduce localized normal forces between the mating box thread crest and pin thread root by 32% in the curved transitions (T1 , T2) of the sections (S1 , S2, S3) in one configuration (compared to just simple angled intersections of straight-line tapers as used in the prior art). This reduction in normal force yields significant reduction in stress reversal magnitudes caused by rotating casing, thereby increasing the endurance limit of the connection assembly (50) for significant increases in allowable rotating time. The actual increase in allowable rotating time (or endurance limit) of the connection assembly (50) is governed by the severity of the wellbore’s dog leg, which defines the stress reversal range. However, in the present configuration used for this illustration in a typical wellbore trajectory, the operator can see an increase in allowable rotating hours of about 30%. In certain situations, this increased rotating time can allow the operator to complete an otherwise unsuccessful event well.

[0101] A secondary benefit of this refined feature is a significant reduction in thread galling potential. So, if a casing string can be recovered, the operator will likely be able to redeploy the casing string, i.e., can re-use all or most of the connection assemblies(50) after cleaning and re-prepping the borehole for rerunning the casing (assuming proper running and retrieval procedures are followed in accordance with manufacturer’s instructions.)

[0102] The multiple tapers (a, 0, - / ) for the sections (S1 , S2, S3) in the coupling 70 reduce the bearing stresses and therefore reduce localized stress risers in the run-out pin threads where assembled connections commonly fail in fatigue under cyclic loading. The continuous internal thread 74 with the curved, tangential transitions (T1 , T2) further reduces localized normal forces and therefore bearing stresses in the run-out pin threads developed during assembly of the connection (i.e., where assembled connections commonly fail in fatigue under cyclic loading).

[0103] The continuous internal thread 75 reduces localized normal forces and therefore bearing stresses by distributing bearing stresses axially over the curved length of the transitions (T1 , T2). This configuration improves fatigue resistance through bearing load distribution over the length of the curved transitions (T1 , T2) and the associated localized, lower bearing stresses distribution in the run-out pin threads (i.e., where assembled connections commonly fail in fatigue under cyclic loading). This continuous internal thread 75 also improves connection fatigue resistance providing longer fatigue life when drilling-with-casing or rotating casing to achieve downhole targets in restricted wellbore conditions.

[0104] Figs. 8A-8B and 9A-9B illustrate schematics based on Finite Element Analysis results from simulation of the disclosed connection assembly 50 in one configuration relative to the prior art. Computer models developed by Computer Assisted Drafting (CAD) software are analyzed using highly sophisticated Finite Element Analysis (FEA) software to produce calculations discussed here. This method is state-of-the-practice for determining various engineering attributes in structural components such as the connections discussed here.

[0105] In particular, Fig. 8A illustrates a schematic analysis of force during mill end makeup of a prior art coupling 20 to the “mill end” of a tubular 10a. This prior art coupling 20 is comparable to that discussed in the background of the present disclosure. By contrast, Fig. 9A illustrates a schematic analysis of force during mill end makeup of the disclosed coupling 70 to the “mill end” of a tubular 60a. As is known, the “mill end” makeup can involve threading the coupling to the “mill end” of a tubular to a hand-tight position in equipment that grips and holds the pipe while another componentof the same equipment grips the coupling and applies torque turning the coupling until a specific make up position is achieved.

[0106] As can be seen in the example of Fig. 8A, the mill end make-up for the prior art coupling 20 produces a calculated normal force (Fn) at T1 (approximately 77,456 lbs). By contrast, the mill end make-up for the disclosed coupling 70 in Fig. 9A produces a lower calculated normal force (Fn) at T1 (approximately 52,636 lbs.) at the same location. The plotted forces show the normal force (Fn) developed on the Box Thread Crest at the center of the transition zone. The numerical values from the normal force (Fn) at this Box Thread Crest are shown for direct comparison. (The models are based on nominal geometry and clearly show significant improvement provided by the disclosed coupling 70 in Fig. 9A). Similar reductions in calculated normal force are also observed at T2 with the present disclosure.

[0107] Meanwhile, Fig. 8B illustrates a schematic analysis of force during field makeup of the “field end” of another tubular 10b to the prior art coupling 20, whereas Fig. 9B illustrates a schematic analysis of force during field makeup of the “field end” of another tubular (e.g., casing) 60b to the disclosed coupling 70. As is known, the “field end” makeup involves handling the casing and stabbing the field end into the open end of the coupling using a stabbing guide, making initial turns during the stabbing process using a chain tong or a power tong in low speed, and threading the “field end” of the tubular (e.g., casing) 60b in the coupling 70. It should be noted here that the mill end can be made up with the disclosed coupling 70 threaded a little past the coupling centerline, which creates more interference on that side of the threaded connection 55. This can help prevent the coupling 70 from turning during makeup of the field end at the rig.

[0108] In this makeup operation, the “lower” tubular 10a of the casing is held with slips in the rig floor’s rotary table while casing tongs grip and turn the “upper” tubular 10b of the casing above the coupling 20 until a power-tight makeup is achieved. Typically, makeup RPMS are initially high (20 to 40 RPM) until the last two full turns, which are then completed in low gear at less than 10 RPMs. Significant torque is applied to produce pin nose engagement between the tubular sections 10a-b, followed by a small increment of axial advancement to lock up or energize the connection. A connection is energized when a significant compression force occurs due to the opposing pin engagement, which causes pre-tensioning about the centerline of the mating coupling 20. Energization of the coupling 20 is necessary to lock the connectionfor better resistance of anticipated downhole loads that may consist of axial (tension and compression), bending, and pressure (internal and external) occurring individually or in any combination. In addition, pin nose contact resists any additional axial advancement of the pin(s) into the coupling 20 that is associated with rotating operations that cause vibration, slip stick, variable friction along the casing string OD, and other factors.

[0109] As can be seen in the example of Fig. 8B, the field end makeup for the prior art coupling 20 produces a mating thread coupling crest to pin root calculated normal force (Fn) (approximately 78,003 lbs) at T1 . By contrast, the field end makeup for the disclosed coupling 70 in Fig. 9B produces a lower calculated normal force (Fn) (approximately 53,262 lbs) at T1 . Again, the plotted forces show the normal force (Fn) developed on the Box Thread Crest at the center of the transition zone. The numerical values from the normal force (Fn) at this Box Thread Crest are shown for direct comparison. (Again, the models are based on nominal geometry and clearly show significant improvement provided by the disclosed coupling 70 in Fig. 9B.) Similar reductions in calculated normal force are also observed at T2 with the present disclosure.

[0110] In previous examples, both of the curved transitions (T1 , T2) have been applied at the intersections (11 , I2) between the sections (S1 , S2, S3). The benefits of the present disclosure can be achieved when the disclosed curved transitions (T1 , T2) have been applied at one or the other of the intersections (11 , I2) between the sections (S1 , S2, S3).

[0111] For example, Fig. 10A illustrates a cross-sectional view of another connection assembly 50 according to the present disclosure, joining two tubular sections 60a-b together. Details of the connection assembly 50 are similar to those discussed above so comparable reference numbers are used for comparable components, even though they may not be described again here. In contrast to previous arrangements, each continuous internal threads 77a-b has a first section (S1 ), a first transition (T1 ), a second section (S2), and a third section (S3). Thus, the curved transition (T1 ) disclosed herein is applied to the first intersection (11) between the first and second sections (S1 ,52), but not at the second intersection (I2) between the second and third sections (S2,53). Instead, the second intersection (I2) between the second and third sections (S2, S3) may have a conventional intersection — e.g., an angle defined between the tapers of the sections (S2, S3). Alternately, in another configuration, the same taper may beused in sections S2 and S3, which eliminates the need for a second intersection 12. Consequently, the configuration would only have two taper sections S1 , S2 with one intersection 11 , thereby eliminating the third section S3.

[0112] Use of the disclosed curved transition (T 1 ) only applied at the first intersection (11) between the first and second sections (S1 , S2) may be suitable for certain sizes, weights (wall thicknesses), and / or grades of the disclosed connection assembly 50. As an alternative, a disclosed curved transition (T2) may alternatively be applied only at the second intersection (I2) between the second and third sections (S2, S3), which may be suitable for other sizes, weights (wall thicknesses), and / or grades of the disclosed connection assembly 50. In any event, heavier wall casing may benefit from applying the disclosed curved transitions (T1 , T2) between both of the adjacent sections (S1 , S2 & S2, S3) as previously described.

[0113] Although the configuration of the coupling 70 may preferably be symmetrical, different combinations of one or both of the intersections (11 , I2) having the disclosed curved transitions (T1 , T2) can be used in the bore 72 at the ends 71a-b of the coupling 70. For example, Fig. 10B illustrates a cross-sectional view of yet another connection assembly 50 according to the present disclosure, joining two tubular sections 60a-b together. Details of the connection assembly 50 are similar to those discussed above so that comparable reference numbers are used for comparable components, even though they may not be described again here. In contrast to previous arrangements, the first continuous internal thread 75a at the field end 71 a of the coupling 70 may have the disclosed curved transitions (T1 , T2) between the sections (S1 , S2, S3) as described herein. However, internal thread 78b at the mill end 71 b of the coupling 70 may have another type of thread, such as a conventional Buttress thread or a sectioned thread as described in the prior art of the background section of the present disclosure. A reverse arrangement is also possible in which the internal thread 75b at the mill end 71 b has the disclosed curved transitions (T1 , T2) between the sections (S1 , S2, S3) as described herein while the internal thread 78a at the field end 71a has a conventional Buttress thread, a sectioned thread as described in the background section of the present disclosure, or another type of thread. This type of arrangement is commonly known in the industry as a cross-over.

[0114] Fig. 11 illustrates a connection assembly 50’ for connecting casing tubulars 60a-b together with a coupling 70 in a slim hole coupled connection according to the present disclosure. As noted in the Background with respect to Figs. 1C-1D, operatorsuse flush and semi-flush integral connections for tight clearance situations between successive casing strings in a well. The connection assembly 50’ disclosed herein can also be used in a slim hole coupled connection for a tight clearance situation between successive casing strings in a well or for a low clearance situation when a casing string is used within the actual wellbore. In contrast to the typical connections, however, the disclosed connection assembly 50’ uses the coupling 70 with a defined smaller than standard OD to connect the pins 64a-b of adjoining casing tubulars 60a-b together.

[0115] The casing tubulars 60a-b have pipe bodies 61 with an internal bore 62. The pipe bodies 61 generally have an outside diameter OD and an inside diameter ID that define a wall thickness Twof the casing tubulars 60a-b. As is customary, the casing tubular 60a-b for the subject applications usually has standard sizes for downhole use with the outside diameters OD ranging from 4 1 / 2 in. to 13 5 / 8 in. Moreover, the casing tubular 60a-b usually has a particular weight (Ibs / ft) defined by the wall thickness Tw suited to the implementation and its requirements. Some example casing sizes include 5 1 / 2 in. and 7 5 / 8 in. The weight for 5 1 / 2 in. casing can range from 11 .5 to 43.10 Ibs / ft, whereas the weight for 7 5 / 8 in. casing can range from 24.0 to 55.30 Ibs / ft. The particular characteristics of casing are defined by American Petroleum Institute (API) casing specifications.

[0116] The pins 64a-b (having thread 65a-b) have an OD turndown 63 that is machined into the outer wall of the casing tubular’s pipe body 61 . As also shown in the detail of Fig. 13, the cylindrical turndown 63 has a reduced diameter Di less than the standard outside diameter OD of the casing tubular 60a-b. The turned-down 63 has a constant diameter initially machined along the entire length Li of the pin 64a-b. Pins 64a-b (having thread 65) is then threaded to a smaller diameter D2 at the nose 66 with an increasing constant taper to the diameter of the turndown 63. Pin thread 65 is a run out thread machined to an extent L2 that lies only partially along the turndown 63 to the nose 66 of the pin 64a-b.

[0117] The coupling 70 is a cylindrical body having an outer diameter D3 and an inner bore 72. As also shown in the detail of Fig. 14, the coupling 70 has a wall thickness Tc defined by the outer diameter D3 and an inner diameter D4 of the coupling’s inner bore 72. The inner bore 72 is threaded with a tapered box thread 75a-b to connect to complimentary tapered pin threads 65a-b on the turned-down pins 64a-b of the adjoining casing tubulars 60a-b. The box thread of continuous internal thread 75a-b has opposing box thread sections along the length L3 of the coupling 70. Inparticular, one box thread section of continuous internal thread 75a generally tapers outward from a center of the coupling’s bore 72 to one of the opposing ends, and the other box thread section 75b generally tapers outward from the center of the coupling’s bore 72 to the other of the opposing ends. Thus, the wall thickness Tc of the coupling 70 constantly varies toward both ends of the coupling 70.

[0118] Although the threads 75a-b of the coupling 70 can each have a single taper, the threads 75a-b of the coupling 70 can have multiple taper sections, such as disclosed in US Pat. Nos. 7,347,459 and 8,075,023, which are incorporated herein by reference. Moreover, Fig. 12 illustrates a schematic view of the coupling 70 for the disclosed slim hole coupled connection including multiple taper sections Si , S2, S3 and including one or more curved transitions T1 , T2 for the thread 75a-b. Three sections are simply shown as an example, and more or fewer sections can be used. The multiple taper sections Si , S2, S3 can reduce contact pressure in areas of the connection assembly 50’, can provide better resistance to long-term, low-level fatigue loading that occurs while rotating casing, and can reduce thread galling common when assembling high interference threaded connections.

[0119] As noted above, the curved transitions T1 , T2 can be curved, tangential transitions defined between the tapered sections Si , S2, S3. These curved transitions T1 , T2 can further minimize the possibility of thread galling and yield significantly increasing connection fatigue resistance by reducing the normal force (or bearing load) imparted between the coupling thread crests and pin thread roots at the transitions T1 and T2 during assembly of the connection improving on the features in the prior art. Moreover, the arrangement having the curved transitions (T1 , T2) effectively converts a point load to a load distributed over a defined length, greatly reducing the applied bearing pressure at the mating interface. The curved transitions (T1 , T2) can provide these and other benefits noted above.

[0120] Returning to Fig. 11 , the pin threads 65a-b thread respectively in the box threads 75a-b at the ends of the coupling 70 to make up the slim hole connection assembly 50’. The bearing surfaces on the noses 66 of the pins 64a-b engage with one another inside the coupling 70. Fig. 11 illustrates the connection assembly 50’ made-up under full power-tight assembly. With power-tight assembly, the bearing faces on the noses 66 of the pins 65a-b meet in the center of the coupling 70 and develop high contact pressure at the mating interface.

[0121] Preferably, compressive forces at the mating noses 66 are carefully controlled during connection assembly. Torque resistance of the connection assembly 50’ is a function of the contact area of the bearing faces on the noses 66 of the pins 65a-b and their material strength. The torque resistance can allow the casing string to be rotated to aid casing string advancement to the intended target. For this reason, the cross- sectional area of the bearing faces on the noses 66 are preferably maximized within the geometric constraints defined by the casing wall thickness Tw, the taper of the pin thread 65a-b, the reduced diameter D2 of the turndown 63, the outside chamfer 67a on the pin nose 66, and the corner break 67b on the inside of the pin nose 66 and the threadform of the thread (65a-b) height, lead, and taper characteristics.

[0122] Considering the cross-sectional area to maximize torque resistance noted above, the discussion now turns to various constraints and variables to make the slim hole connection assembly 50’ suited for a particular implementation. As expected, the coupling 70 has an outer diameter D3 that can be selected to meet regulatory clearance requirements. For example, the outer diameter D3 can be greater than the tubular’s outer diameter OD by a predetermined amount that would allow the connection assembly 50’ to maintain a required annular clearance with a surrounding casing or open hole wellbore wall in which the connection assembly 50’ is deployed. In fact, the outside diameter D3 of this coupled connection assembly 50’ can be similar to the profile for a semi-flush, integral joint connection (11 B: Fig. 1C) and can accomplish the same clearance objectives. Thus, the actual outer diameter Dsof the coupling 70 is usually a set constraint for the implementation to meet minimum regulatory annular clearance requirements.A. MAXIMIZING FOR PRESSURE AND TENSION EFFICIENCIES

[0123] According to one embodiment, the connection assembly 50’ of the present disclosure is configured so that it is optimized for both pressure and tension efficiencies (with the torque resistance added). To do this, the coupled connection assembly 50’ of the present disclosure is designed so that critical cross-sections of the pipe 60a-b are generally balanced to the critical cross-sections of the coupling 70. This results in a matched-strength design as understood in the art. In particular, the basic concept is that if the relative cross-sectional areas (which are related to the individual thicknesses) of the mating parts are about the same, the respective performance properties of the threaded connection 55 relative to the pipe body 61 will be matched. This is often referred to as a “matched strength” design. When the geometry does not allow for atruly “matched strength” design, the goal then becomes utilizing the available cross- sectional area(s) to achieve a target minimum level of performance in the threaded connection 55 relative to the pipe body 61 . This is often referred to as connection efficiency. As seen in the above discussions, a limited clearance situation having an annular clearance limit calls for reduced outside diameter of the coupling 70 at the threaded connection 55. In this design, the pipe 60a-b has a smaller secondary OD at the turndown 63 for threading the pins 64a-b. This smaller OD at the turndown 63 allows a portion of the pin’s cross-section to be shared with the coupling 70, which increases the cross-section of the coupling 70 and makes the coupling 70 stronger to meet the target performance efficiencies. In these cases, the target efficiency of the present disclosure is a minimum of 70%.

[0124] As shown in Figs. 11 and 13, the casing or pipe 60a-b has an outside diameter cylinder turndown 63, which is formed on the pipe body 61 prior to machining the pin thread 65 on the pin 64. In Fig. 13, the pipe 60a-b is schematically shown with critical cross-sections. The critical cross-section for the pipe body 61 is depicted as CPB; the critical cross-section for the cylinder turndown 63 is depicted as CPT; and the critical cross-section for the pin is depicted as CP. In Fig. 14, the coupling 70 is schematically shown with critical cross-sections. The critical cross-section for the coupling tension is depicted as CCT; and the critical cross-section for the coupling pressure is depicted as CCP.

[0125] As noted herein, this connection assembly 50’ uses the coupling 70 for extreme clearance situations. The turndown 63 of the pipe 60a-b and the threading on the pin 64 are married to complementary features of the coupling 70 in a design process that balances all of the geometric features to yield one or more targeted connection efficiencies, such as yield strength, pressure, tension, and / or torque resistance. Notably, the critical cross-sections that are balanced to produce the targeted connection efficiencies for tension and pressure in each member define four specific and different locations. Further discussion of these critical cross-sections is described below.

[0126] Looking generally at balancing the critical cross-sections to maximize the pressure and tension efficiencies of the connection assembly 50’, it is understood that the pin threads 65a-b on the tubulars’ pins 64a-b screw into the mating female box threads 75a-b of the coupling 70 to form the slim hole connection assembly 50’. Given that the outside diameter D3 for the coupling 70 is a set constraint dictated by theallowable annular clearance for the implementation, the mating cross-sections CCT, CCP of the connection assembly 50’ are configured / balanced to produce a connection assembly 50’ that meets a certain set of performance requirements, such as tension, and pressure requirements. In other words, because the outside diameter and wall thickness of the casing 60a-b and the connection’s outside diameter (i.e., the coupling’s outside diameter D3) are defined by the casing string geometry, the slim hole connection assembly 50’ is configured to generally balance the relative critical crosssections CCP, CPT of the coupling 70 and the pins 64a-b to (a) achieve a generally balanced strength relative to one another, and (b) meet an efficiency target with respect to the pipe body tension and pressure resistances.

[0127] Looking with more particularity at balancing the critical cross-sections, several factors are involved in configuring the slim hole connection assembly 50’ as the relative strength of the connected features must be calculated and considered. The connected features under consideration include the pipe body 61 , the cylinder turndown 63, the threaded pins 64a-b, and the coupling 70. Each of these features is rated for tension and pressure resistance. In addition, each of these features has a critical cross-section that defines its individual tension and pressure resistance. Moreover, the location of the critical cross-sections in the pins 64a-b and the coupling 70 are different for tension and pressure resistance. In the design of the slim hole connections assembly 50’, a series of properties are determined for each of these features of the slim hole connection assembly 50’ comprised of the pipe body 61 , the turn down 63, the threaded pins 64a-b, and the coupling 70. A series of iterative calculations involving multiple strength equations can be used to yield the relative four critical cross-sections of these features to achieve a target connection efficiency relative to the performance properties of the pipe body 61.

[0128] Industry-standard equations defined by API can be used to calculate the individual uni-axial performance properties of each of the features at the prescribed locations of the critical cross-sections. The API equations are approximations only as they are derived for similar connection geometry. However, the modifications proposed herein for the connection assembly 50’ are sufficiently different for some properties that the API equations may be overly conservative or may not directly apply.

[0129] For this reason, more rigorous analytical methods, such as Finite Element Analysis, can also be used to evaluate connection performance in terms of stresses and strains when multiple loads are applied in combination. These rigorous analyticalmethods can more closely simulate the real-world conditions by considering various possible applied load combinations when casing and connections are deployed and throughout the intended service life of a well. Considering the noted critical crosssections for the features and using the computational analysis, the design of the present disclosure seeks to achieve a balance among the various relative cross-sections (CPB for the pipe body 61 , CPT for the cylinder turndown 63, CP for the pins 64a-b, CCT for the coupling tension, and CCP for the coupling pressure) to achieve target efficiencies of the coupled connection assembly 50’ relative to the pipe body 61.

[0130] Torque resistance is another important feature of the subject design because operators often have to rotate the casing to break the downhole skin friction of the casing string to help advance the casing string to the target depth when deploying the string in the well. Torque is achieved by maximizing the bearing areas on the pin noses 66 of the pins 64a-b. This is necessarily constrained by the wall thickness of the pipe body 61 , the diameter Di of the cylinder turndown 63, and the threadform height geometry and taper of the threads 65a-b on the pipe ends 64a-b, which ultimately define the dimensions for the bearing face on the pin nose 66. Torque resistance is determined by the area of the threadform load flanks of the thread 65a-b that drive the pin noses 66 together as they are rotationally advanced into the coupling 70, which ultimately forms the connection 60a-b. While torque is not an industry-specified performance property, most operators are vitally interested in the torque capabilities of casing connections because high torque resistance provides valuable utility in low clearance casing strings and well construction where deploying casing to target can be challenging. In some instances, the ability to apply torque to rotate the casing can make the difference between completing a well or having to abandon efforts due to stuck pipe or a lost-hole situation at a potential cost in the millions of dollars. The prior art connections 11A-B in Figs. 1C-1D do not have significant torque resistance, and therefore they cannot be rotated if needed to achieve a target depth.

[0131] As is usual, parameters of the casing tubular 60a-b, such as the casing outer diameter OD, casing weight, wall thickness Tw, yield strength, ultimate strength, modulus of elasticity, and the like, are designed for the particular implementation for which the casing tubular 60a-b is to be used. Project-specific well requirements ultimately indicate the diameter, wall thickness, and performance properties of the casing tubular 60a-b needed for any given well.

[0132] Other parameters of the connection assembly 50’ are defined by the geometry of the implementation. Again, the coupling’s outside diameter D3 is defined by the required annular clearance for the casing tubular’s size being used inside a surrounding casing or open hole wellbore wall. In general, the nominal wall thickness Tcof the coupling 70 is a function of the annular clearance requirements, targeted performance properties, and diameter Di of the turn down 63 required for the application. It may be possible that a configuration of the present disclosure may not be able to achieve a target performance property due to the casing wall thickness Tw. In these instances, the only option for the operator may be to use a flush connection (which will also likely be insufficient) or a semi-flush premium connection as a default and to abandon any additional utility offered by a connection that can be rotated. In extreme cases, the entire casing string program may have to be redesigned to fit wellspecific criteria unachievable by the slim-hole configuration disclosed here and / or flush and semi-flush connections available in the industry.

[0133] The thread forms used for the pin and box threads 65, 75 can comport with industry-standard thread forms as to thread taper, thread height, API Buttress Connection (BC) perfect thread length, thread pitch, and the like. Proprietary thread designs can also be used for the pin and box threads 65, 75 but this would require threadform specific equations for design optimization. The use of industry-standard threads allows for easier and lower-cost access to accessories needed for field deployment. The pin and box threads 65a-b, 75a-b preferably use an interference fit thread designed for tension and pressure resistance. Such thread interference fit connections are often referred to as “semi-premium” connections, which are commonly used for challenging as opposed to the most critical well applications.

[0134] Ultimately, given these constraints, the connection assembly 50’ is designed to meet a desired set of strength ratings (performance properties). For example, the connection assembly 50’ preferably has desirable tension efficiency compared to the pipe body 61 of the tubular 60a-b. An example of a typical target efficiency is 70%, but a higher or even lower efficiency may be acceptable depending on the well-specific application. In other words, the minimum yield for the connection assembly 50’ at the cylinder turn down 63’s diameter Di , at the pin threads 64a-b, and at the coupling 70 can be designed to be at least 70% of the minimum yield of the pipe body 61 of the casing tubular 60a-b. It is also usually desirable for the connection assembly 50’pressure efficiency to be at least 70% of pipe body 61. Other values for the strength rating can be used depending on the requirements of a given implementation.

[0135] Taking into consideration all the above constraints for the casing 60a-b and coupling 70, features of the connection assembly 50’ are calculated so the desired strength ratings or connection efficiencies can be achieved. These features include values for the turndown diameter Di , the taper of the pin 64a-b from the turndown diameter D1 to the end diameter D2, and the cross-sectional area of the pin nose 66. These variables define the critical cross-section CpTfor the turndown 63 and the critical cross-section CP for the pin 64. Further, the outside diameter D3 of the coupling 70 is defined by the regulatory clearance requirements taking into account the inside diameter of the outer string or open hole wellbore ID. The balancing of critical crosssections determines the critical tension cross-section CCT and the critical pressure cross-section CCP of the coupling 70.

[0136] Values for these features provide calculations for the critical thickness at the turndown 63, the bearing area of the nominal pin nose 66 of the pin 64 after the turndown 63, a nominal force that yields the pin nose 66, yield torque after the turndown 63, pipe thread strength after the turndown 63, minimum internal pressure rating for the pin 64 after the turndown 63, and the like.

[0137] A specific example can further the discussion of balanced cross- sections / matched-strength design discussed above. In the specific example, a tubular (e.g., casing) 60 having an outer diameter OD of 5-1 / 2 in., a weight of 20 Ibs / ft, and a wall thickness Twof 0.361 in. will be used to illustrate the connection tension rating. Although this discussion will focus on the strength through geometry principle as it relies on cross-sectional areas to determine tensile ratings, the cross-sectional area is also related to other performance properties, including, but not limited to, pressure, fatigue, collapse, bending, and buckling. As will therefore be appreciated with the benefit of the present disclosure, an interactive process is used to optimize the design for more than just tension. Hence, the design becomes iterative as described previously to yield the best design for all requisite performance properties.

[0138] In designing a well, the drilling engineer determines if the tensile rating of each casing component (casing, coupling, and thread strength) is sufficient to resist the anticipated downhole service loads after applying an appropriate factor of safety.Tensile capacity in the simplest form is the axial load that can be resisted by a component, say the casing body. This load is calculated based on the material’s yieldstrength and the casing body’s cross-sectional area. Thus, the area of the casing body is the engineering critical cross-section due to the fact that once a tensile force exceeds the calculated tensile capacity, material yielding will occur followed by failure of the part (if the load is not reduced or removed) across the cross-section of the casing body.With respect to tension (for example), tensile ratings for each component of the connection are compared. The lowest tensile rating among these components defines the limiting property that will be considered in the string design.

[0139] As discussed, “balancing critical cross-sections” is the means for getting the thickness of the coupling 70 to have a thickness similar to the pipe body 61 . While they cannot realistically have the same thickness, the respective cross-sectional areas can get close enough to achieve the design objectives. A smaller cross-sectional thickness on the larger diameter, outer member (e.g., coupling 70) can have a greater cross- sectional area than a thicker section on a pipe body 61 , which has a smaller outside diameter. This will become evident later in the discussion.

[0140] One annular clearance requirement is given by the BUREAU OF LAND MANAGEMENT 43 CFR 3160, which stipulates that “casing collars shall have a minimum clearance of 0.422 inches on all sides in the hole / casing annulus, with recognition that variances can be granted for justified exceptions” (Federal Register Vol. 53, No. 223; November 18, 1988). This equals 0.844 in. when considered a diametrical annular clearance.

[0141] For the example having 5-1 / 2” OD casing, such a requirement sets the allowable coupling OD at 5.900 in., which is the OD constraint, assuming standard bit sizes typically used in drilling out of 7 5 / 8-in. OD casing. The casing size, weight, wall thickness, and grade suitable for the intended service sets the casing’s inside diameter (ID), which is the ID constraint. The difference between the coupling’s OD and the casing’s ID defines the overall cross-sectional area of the connection or the structural real estate a designer has available to meet the downhole loading conditions with acceptable factors of safety.

[0142] Fig. 15A shows a configuration having a standard casing 10 with 5-1 / 2” OD, 20.00 ppf (0.361” wall). An example coupling 20a is shown with a standard OD of 6.300” as represented by the dashed line. Another example coupling 20b is also shown with the allowable coupling OD (5.900”), which would be ideal if reducing a standard coupling’s outer diameter was the simple solution to meet the minimum allowableclearance. Unfortunately, such a simple reduction in the coupling’s outer diameter is not the best solution.

[0143] The following discussion provides particular dimensional and tension ratings for the two scenarios shown in Fig. 15A. In the first scenario for the coupling 20a, the coupling OD is 6.300”. The cross-sectional area of the pipe body 13 is 5.8282 sq. in. The plain end yield strength of the pipe body 13 is 641.1 kips (tension rating per API). The cross-sectional area of the coupling 20a with the 6.300 OD is 8.5267 sq. in. The minimum tension of the coupling 20a with the 6.300 OD is 1 ,012.5 kips (per API). The pipe thread strength is 667 kips (per API), the tensile efficiency of the coupling 20a is greater than 100%, and the tensile efficiency of the connection is greater than 88.1 %.

[0144] In this first scenario, the threaded connection 25 is clearly pipe weak in tension because (i) the area of the pipe body 13 is much smaller than the area of the 6.300” OD coupling 20a (i.e., 8.5268 vs 5.8282 sq. in.) and (ii) the tension rating of the coupling 20a is greater than the pipe body 13 (i.e., 1 ,012.5 vs 641.1 kips).

[0145] In the second scenario of the coupling 20b, the coupling OD is 5.900 in. The area of the pipe body 13 is 5.8282 sq. in. The plain end yield strength of the pipe body 13 is 641 .1 kips (tension rating per API). The area of the coupling 20b with the 5.900 OD is 4.6939 sq. in. The minimum tension of the coupling 20b with the 5.900 OD is 557.4 kips (per API), and the tensile efficiency of the coupling 20b is 86.9%.

[0146] In this second scenario, the threaded connection 25 is clearly coupling weak in tension because (i) the area of the 5.900” OD coupling 20b is less than the area of the pipe body 13 (4.6939 vs 5.8282 sq. in.) and (ii) the tension rating of the coupling 20b is less than the pipe body 13 (491.5 vs 641.1 kips). As seen, this coupling 20 with the 5.900” OD using a standard pin thread fails to meet the design requirements for this casing 10.

[0147] The teachings of the present disclosure seek to solve this problem. Turning now to Fig. 15B, a configuration according to the present disclosure is shown having a casing 60 with 5-1 / 2” OD, 20.00 ppf (0.361” wall) and having a coupling 70. The casing 60 has a turndown 63 at the end of the pipe body 61 , and the pipe body 61 has been machined with a commensurately smaller pin thread 65. A portion of the pin member has been taken and shifted to the coupling 70. In this example, the casing 60 has a new cross-section with an OD’ of 5.395 in.

[0148] Following the calculation methodology as presented above, a single configuration is now discussed for illustrative purposes only. This example presents anoptimized design providing a solution for 5-1 / 2 in OD, 20.00 ppf (0.361”) wall being deployed inside of 7-5 / 8” OD casing even though this example addresses only the tension rating and efficiencies.

[0149] The techniques of the present disclosure seek to balance the critical cross- sectional areas of the coupling 70 and pipe thread 65 to achieve targeted performance property efficiencies. Calculating the joint strength for the threaded connection 55 for this connection assembly 50’ requires determination of the minimum tension rating for the coupling 70, pipe thread 65, and the turndown 63 of the pin 64. The results of three (3) equations are solved to determine the tensile rating (minimum of the three calculated values) of the threaded connection 55.

[0150] In the third scenario of the present configuration in Fig. 15B, the coupling 70 has an OD of 5.900 in, and the pipe OD has the turndown 63 to 5.395 in. at the pin 64. The cross-sectional area of the pipe body 61 is 5.8282 sq. in., and the plain end yield strength of the pipe body 61 is 641 .1 kips (tension rating per API). The cross-sectional area of the coupling 70 with the 5.900 OD is 5.5709 sq. in., and the minimum tension of the coupling 70 with the 5.900-in. OD is 661 .5 kips (per API). The tensile efficiency of the coupling 70 is greater than 100%.

[0151] The present configuration in the third scenario provides a better solution by increasing the cross-section area of the coupling 70 to 5.5719 sq. in. from 4.6939 sq. (18.7%) and by providing a tensile efficiency of the coupling 70 to over 100% compared with only 86.9% for the case illustrated in the second scenario. This new configuration is pipe weak, which is the preferred situation within the industry. It should be pointed out that Figs. 15A-15B are not to scale. Although not visually obvious, the cross-section of the coupling 70 in Fig. 15B is bigger (and therefore stronger) than the coupling 20 in Fig. 15A. As can be seen, the turndown 63 of the casing 60 allows the available cross- sectional real estate to be shared with the coupling 70, thereby bolstering the coupling’s strength.

[0152] If the couplings’ pressure ratings are compared between the second and third scenarios, there is even greater improvement in connection efficiency. In the second scenario of Fig. 15A with coupling 20b, the minimum internal pressure for the 5.900-in. OD coupling 20b is 9,887 psi, and the pressure efficiency is 78%. In the third scenario of Fig. 15B, the minimum internal pressure for the 5.900-in. OD coupling 70 is 11 ,845 psi, and the pressure efficiency is 93.7%.

[0153] Based on this example, the balance of cross-sectional thicknesses (or areas) of the mating members of the pin 64 and coupling 70 increases the geometric strength by best utilizing the material between the allowable OD of the coupling 70 set by regulation and the ID of the pipe body 61 set by the selected casing size and weight. Accordingly, the teachings of the present disclosure seek to optimize performance property ratings through cross-sectional balance improvements that ultimately provide performance rating enhancements while always striving for a matched-strength design.

[0154] Retuning to Fig. 11 , the example casing tubular 60a-b having an OD of 5-1 / 2 in., a weight of 20 Ibs / ft, and a wall thickness Twof 0.361 in. can be used with a coupling 70 having an outside diameter D3 of 5.900 in. to meet the annular space requirements. In one example configuration similar to that discussed above with respect to Fig. 15B, the features can be designed with the nominal turndown diameter Di of about 5.375 in., the cross-sectional area CPof pipe end 64 after the turndown 63 can be about 5.398 in2, and other calculated values. The pipe body 61 in this example with a minimum yield strength of about 125,000 psi would have a plain end tension rating of about 728.5 kips. The present connection assembly 50’ machined from the same material with the minimum yield strength of about 125,000 psi would then be designed to produce a minimum coupling tension rating of about 607.3 kips (i.e., at least 80% of the pipe body’s tension rating although other percentage values could be targeted).

[0155] Oilfield operators need connections with high torque ratings for rotating casing to assist target achievement in long lateral wells. The connection assembly 50’ disclosed herein provides this attractive utility through engagement between opposing pin noses 66 on the casing tubulars 60a-b by maximizing the nose bearing area on the pin noses 66 to provide torque resistance. The connection assembly 50’ of the present disclosure can be produced at a lower cost than flush or semi-flush connections, and the field deployment can also be easier.

[0156] In general, there may be standard designs for the connection assembly 50’ disclosed herein, but there will only be one standard design for each casing OD and wall thickness. In one standard design, the connection assembly 50’ is optimized to meet about 70 to 80% tension and pressure efficiency of a base pipe body while providing a maximum torque resistance. In these cases, torque will be maximized given constraints on the available area for the pin nose 66.

[0157] However, the desired efficiency of the connection assembly 50’ can vary based on the application. In some cases, the tension efficiency of the connectionassembly 50’ may be the desired factor. In other cases, the efficiencies for pressure resistance or torque for the connection may be the desired factor. Accordingly, the connection assembly 50’ may comprise several variations for a single size, weight, and grade of casing to suit different applications. Overall, there are a number of variables so achieving the desired connection assembly 50’ requires multiple iterations depending on what target efficiencies are desired.B. MAXIMIZING FOR PRESSURE EFFICIENCY

[0158] The discussion provided above focused on configuring the connection assembly 50’ so that it optimizes for both pressure and tension efficiencies (with the torque resistance added). A connection assembly 50’ can be configured according to the present disclosure to maximize mainly the pressure efficiency of the connection assembly 50’. The critical pressure rating that defines the efficiency of the connection assembly 50’ is the lower of the pressure rating for the cylinder turndown 63, the minimum internal pressure for the coupling 70, and the connection leak resistance. Values for each of these may be defined by API formulas.

[0159] To maximize the connection assembly 50’ mainly for pressure efficiency, for example, the coupling 70 can be designed so that the critical coupling cross-section CCP is as large as possible relative to the pipe body cross-section CPB. Pressure is a function of the pipe body cross-section CPB, the turndown cross-section CPT, and the critical coupling cross-section CCP. Accordingly, the design objective is to balance these cross-sections (CPT and CCP) within the constraints of: (i) the pipe body wall thickness Tw and (ii) the limiting coupling OD defined by regulations. Other methods can be used to meet the target requirements that might include changing the thread taper, height, and length.C. MAXIMIZING TENSION EFFICIENCY

[0160] The discussion provided above focused on configuring the connection assembly 50’ so that it optimizes for both pressure and tension efficiencies, or it maximizes mainly the pressure efficiency of the connection assembly 50’. A connection assembly 50’ can be configured according to the present disclosure to maximize mainly the tension efficiency of the connection assembly 50’. In general, the critical tension rating that defines the connection’s tension efficiency is the lower of the tension rating for the cylinder turndown 63, the pipe thread strength, and the minimum coupling tension. Values for each of these may be defined by established API formulas (API Technical Report TR5C3).

[0161] There are several options to maximize the connection assembly 50’ mainly for tension efficiency, for example, that ultimately lead to maximizing the ratio of cross- sectional areas CCT to CPB. One method can increase the cylinder turn diameter D1 to provide a larger critical cross-section CCT in the coupling 70. However, given the constraints for the OD of the coupling 70 and the wall thickness Twfor the pipe body, other design techniques such as changing the thread taper, thread height, or overall length may be required to meet the target performance requirements of connection assembly 50’.D. DETAILS OF TURNDOWN GEOMETRY

[0162] Although the turndown 63 helps the connection assembly 50’ fit within a particular clearance requirement as noted, the features of the turndown 63 (e.g., its geometry, diameter, taper, placement on pipe end, etc.) are designed in an iterative process based on several considerations to achieve the targeted efficiencies noted herein.

[0163] In particular, a length of the turndown 63 can be configured to avoid allowing extra thread interference from occurring, which can cause corresponding high hoop stresses in the coupling 70 at the ends. The transition from the cylinder turn diameter Di out to the pipe body’s OD should be at a shallow angle with a fillet transition. An internal fillet avoids creating a stress riser at the corner that could be detrimental to downhole performance of the connection assembly 50’ under anticipated service load conditions.

[0164] Preferably, the starting thread 75a-b in the coupling 70 does not radially interfere in any way with the transition from the cylinder turndown’s diameter Di to the pipe body’s OD. Because the wall thickness Tw of the pipe body (61 ) is greater than the wall thickness CTW after cylinder turndown 63, the thicker part is stronger. Thus, it is preferred that the transition occurs a sufficient distance away from the ends of the coupling 70 after power tight assembly to avoid imparting additional hoop stresses in the coupling 70 that could detract from downhole performance under service loading conditions.E. DETAILS OF MATING THREADS

[0165] Discussion now turns to details of the mating threads 65a-b, 75a-b, as referenced in Fig. 16. The pins 64a-b and the coupling 70 preferably use high radial interference mating threads 65a-b, 75a-b. During assembly, the box ends of the coupling 70 are forced to radially expand outward (E). At the same time, the pins 64a-bare forced to compress radially inward (C). The assembly is synonymous with forcing a large external conical member into a smaller internal conical member. The amount of outward box expansion (E) of the coupling 70 and inward pin compression (C) of the pins 64a-b for each member is relative to their individual cross-sectional areas, i.e., their relative strengths. It should be noted that the cross-sections of each member vary along their length which further complicates the problem.

[0166] Given that both members 64a-b, 70 are tapered, the cross-sections constantly vary such that the thinnest portion of the pin64a-b (i.e., the nose 66) ends up adjacent to the thickest part of the coupling 70 at its midpoint. Conversely, the thickest part of the pins 64a-b ends up adjacent to the thinnest part of the coupling 70. The resulting stresses form a 3-dimensionsal effect that includes (1 ) radial expansion / contraction of the two parts combined with (2) shortening of the coupling 70 as the coupling 70 expands outward (E), and (3) lengthening the pins 64a-b as the pins 64a-b are compressed inward (C). The shortening of the coupling 70 as it expands outward and the lengthening of the pin 64a-b as it is compressed inward is known as the Poisson effect in the art of engineering. In addition, there is also axial bending of the members 64a-b, 70 that occurs with the connection assembly 50 connected to the power tight position. To illustrate the effect, hoop stress and displacements are schematically shown in Fig. 16 for the connection assembly 50’ after assembly. The scale of the displacements is exaggerated to highlight the changes. As expected, the relative displacements of the members 64a-b, 70 cause strain and therefore stress.

[0167] Large hoop stress in couplings or box members that are made of susceptible materials and are placed in unsuitable environments can cause a common industry problem referred to as “environmentally-assisted cracking.” In certain environments, for example, high-stress areas of the members create pockets for free hydrogen to reside. As free hydrogen increases and crowds into the interstitial spaces (i.e., the fabric of the steel), cracks develop causing failure of the box member. These failures occur catastrophically and without notice. They also occur randomly within various strings of adjacent wells. The phenomenon is referred to as “environmentally-assisted cracking” because in many cases the source of the free hydrogen cannot be determined and can develop quickly or over some time before failure occurs. The typical mode of failure is hydrogen or corrosion-related embrittlement.

[0168] According to the present disclosure, the high radial interference mating threads 65a-b, 75a-b and the mating cross-sectional areas of the members 64a-b, 70are modeled to determine the stress levels and displacements that can develop during assembly of the connection assembly 50’. In this way, the disclosed techniques for balancing the mating cross-sections described herein to optimize the connection assembly 50’ for pressure and / or tension efficiencies can also be configured to yield low hoop stresses at the connection assembly 50’. Attention to avoiding high hoop stress levels can be performed to reduce the risk of “environmentally-assisted cracking” when the casing is to be made of susceptible materials to meet required performance properties. Additionally, configuring the connection assembly 50’ to have low hoop stress levels after assembly can leave more residual strength to resist anticipated downhole load conditions that are additive to the made-up stress states of the connection assembly 50’.F. DETAILS OF PIN NOSE GEOMETRY

[0169] Although the pin noses 66 of the connection assembly 50’ engage with one another and create torque resistance, the features of the pin nose 66 (e.g., geometry, chamfer, corner break, thread root clearance, etc.) are designed to have several beneficial characteristics.

[0170] Several factors impact the final geometry of the bearing face on the pin nose 66. In particular, threadform characteristics, such as thread height, taper, and length, can have an impact on the final geometry. The wall thickness Tw of the pipe body 61 also constrains the bearing area available on the face of the pin nose 66. The pin nose 66 also has an OD chamfer 67a, which can aid in stabbing the pin member into the box member of the connection assembly 50’.

[0171] When considering the threadform of the thread 65, the thread’s root ideally exits on the chamfer 67a of the pin nose 66 and not at the bearing face of the pin nose 66. Otherwise, the threadform of the thread 65 would reduce the bearing area of the pin nose 66 and could compromise the torque resistance of the connection assembly 50’.

[0172] Lastly, an ID chamfer can be provided on the pin nose 66 as is customary to eliminate a sharp corner that could cause injury to a careless worker. In the industry, there are no standardized specifications for such an ID chamfer. Most manufacturers prefer to create this ID chamfer during threading for efficiency and corresponding low cost. Given all of the variables associated with casings produced by mills around the world, machine programmers typically create an aggressive diameter for an ID chamfer that is assured to eliminate the aforementioned sharp corner. However, this practicesignificantly reduces the bearing area of the pin nose 66 and therefore can reduce the torque capacity of the connection 60a-b.

[0173] In the present configuration, the pin nose 66 is preferably treated with an ID corner break 67b instead of an ID chamfer. This ID corner break 67b is effectively a minimal chamfer that mitigates the issue of sharp corner safety, but maximizes the bearing area for the pin nose 66 for added torque capacity. In most cases, this ID corner break 67b can be produced in a manual process to avoid reducing the bearing face of the pin nose 66.G. SUMMARY

[0174] In summary, the tubular systems disclosed above with reference to Figs. 2A through 15 can be used for drilling a well and for casing off an open hole. In particular, the tubular system can be used as a drillstring and can have a drill bit attached to its bottom end. Additionally, used in a full casing program, the disclosed tubular system can be used in implementations consisting of multiple telescoping strings of casing in which a minimum annular clearance is required between the successive casing strings deployed in the well. The tubular systems disclosed give operators significant advantage in difficult well conditions, allowing application of high torque to rotate casing to assist achieving a target setting depth or to free stuck pipe. Without these important features, operators have limited options and may often need to abandon a well in which potentially millions of dollars have already been expended.

[0175] As already noted, standard API threaded connections have a very loose tolerance allowance, with the allowable make-up variation being 0.575 inches (i.e., 2.88 turns) between the minimum and maximum make-up positions. The threaded connections 55 of the present disclosure used in the more challenging implementations for drilling with casing and for fitting within minimum annular clearances cannot provide the best performance if such a loose allowance is tolerated. Instead, the threaded connections 55 of the present disclosure can have an allowable make-up variation that is 0.125 inch (0.56 turns) between the minimum and maximum make-up positions. In this way, the threaded connections 55 of the present disclosure can offer tighter diametric tolerances on the mating pin and coupling threads, which can provide improved connection performance for the more challenging implementations noted herein.

[0176] As disclosed herein, a number of variables and equations are required to find an optimum configuration to meet target performance ratings for given geometryconstraints (e.g., pipe inside diameter, coupling outside diameter, minimum annular clearance, etc.). The solution requires solving several equations using variables in a complicated iterative procedure to achieve the target performance rating(s). The novel approach according to the disclosed techniques shares the geometry of the pin member with the coupling member by accounting for how the relevant equations are interrelated with one another. With each iteration there is always a tradeoff. On factor (performance property rating) may improve while another is reduced. Consequently, the process takes multiple iterations to solve the suite of equations that finishes with an acceptable configuration for the implementation at hand.

[0177] Configurations of the present disclosure can be characterized by the following clauses:1 . A coupling for joining tubulars, the coupling comprising: a body (70) having a first end (71a) and a second end (71 b) and defining a bore (72) therethrough; at least a first portion of the bore (72) defining a first continuous internal thread (75a), the first continuous internal thread (75a) having a first section (S1 ), a second section (S2), and a third section (S3), the first section (S1 ) disposed at the first end (71a), the second section (S2) connected with the first section (S1 ) at a first intersection (11), the third section (S3) connected with the second section (S2) at a second intersection (I2), an internal diameter (76a) of the first continuous internal thread (75a) for the first section (S1 ) converging linearly inward toward the bore (72) at a first angle (a) from a first point (p1 ) to the first intersection (11 ), the internal diameter (76a) for the second section (S2) converging linearly inward toward the bore (72) at a second angle (P) from the first intersection (11 ) to the second intersection (I2), the second angle (P) being less than the first angle (a), the internal diameter (76a) for the third section (S3) diverging linearly outward from the bore (72) at a third angle (- / ) from the second intersection (I2) to a second point (p2), wherein the internal diameter of the first continuous internal thread (75a) defines at least one of: (i) a first curvature (T1 ) transitioning between the first and second angles (a, P) at the first intersection (11 ) of the first and second sections (S1 , S2), and (ii) a second curvature (T2) transitioning between the second and third angles (P , - / ) at the second intersection (I2) of the second and third sections (S2, S3).For example, the internal diameter (76a) of the first continuous internal thread (75a) at at least one of (i) the first intersection (11 ) and (ii) the second intersection (I2) can define a tangential curvature entering and exiting transitions from a respective one of: (i) the first angle (a) to the second angle (P), and (ii) the second angle (P) to the third angle (- / ).2. The coupling of clause 1 , wherein the respective at least one of: (i) the first curvature (T1 ) is tangential to the first and second angles (a, P); and (ii) the second curvature (T2) is tangential to the second and third angles (P, - / ). In other words, for example, the respective one of (i) the first angle (a) and the second angle (P) and (ii) the second angle (P) and the third angle (- / ) are tangential to the curvature.3. The coupling of clause 1 or 2, wherein: the internal diameter (76a) defines the first curvature (T1 ) at the first intersection (H ); the internal diameter (76a) for the first section (S1 ) converging linearly inward toward the bore (72) at the first angle (a) from the first point (p1) to the first curvature (T1 ); the internal diameter (76a) for the second section (S2) converging linearly inward toward the bore (72) at the second angle (P) from the first curvature (T1 ) to the second intersection (I2); and the first curvature (T1 ) is tangential to the first angle (a) to the second angle (P).4. The coupling of clause 1 , 2 or 3, wherein: the internal diameter (76a) defines the second curvature (T2) at the second intersection (I2); the internal diameter (76a) for the second section (S2) converging linearly inward toward the bore (72) at the second angle (P) from the first intersection (11) to the second curvature (T2); the internal diameter (76a) for the third section (S3) diverging linearly outward from the bore (72) at the third angle (- / ) from the second curvature (T2) to the second point; and the second curvature (T2) is tangential to the second angle (P) to the third angle (- / )■5. The coupling of clause 4, wherein the first curvature (T1 ) is tangential to the first angle (a) and the second angle (P); and wherein the second curvature (T2) is tangential to the second angle (P) and the third angle (- / ).6. The coupling of any one of clauses 1 to 5, wherein the first continuous internal thread (75a) is configured to mate with a pin (64a) of one of the tubulars (60a); and wherein: the pin (64a) having an external buttress-type thread (65a); or the pin (64a) having a pin thread (65a) with a single taper angle (P’)> and at least one of the first angle (a), the second angle (P), and the third angle (- / ) of the first continuous internal thread (75a) is approximate (i.e., matches within machining tolerances) to the single taper angle (P’) of the pin thread (65a), optionally wherein the second angle (P) of the second section (S2) is approximate (i.e., matches within machining tolerances) to the single taper angle (P’) of the pin thread (65a) on the pin (64a) of the tubular (60a).7. The coupling of any one of clauses 1 to 9, wherein the at least one of the first and second curvatures (T1 , T2) spans a helical profile around a portion of the respective continuous internal thread where one of the sections transitions to the adjacent section.8. The coupling of clause 7, wherein: the helical profile, of the first continuous internal thread (75a) being formed by a threading tool, spans between a half (T ) rotation to one (1 ) full rotation of the coupling as the coupling is threaded by the threading tool; the first continuous internal thread (75a) has a thread pitch, and an axial length (a of the helical profile is between 1 / 4 to 1 / 2 of the thread pitch to 1 / 2 to 1 of the thread pitch; or an extent of the helical profile is defined by (i) a relative thickness of the body (70) and the tubular (60a-b) and (ii) radii of the first and second curvatures (T 1 , T2).For example, the at least one of the first and second curvatures (T 1 , T2) spans a helical profile around a portion of the respective continuous internal thread between (i) a half (>2) rotation of the coupling as the coupling is threaded by a threading tool (or an axial length of 1 / 4 to 1 / 2 of the thread pitch) and (ii) one (1) full rotation of the coupling as the coupling is threaded by a threading tool (or an axial length of 1 / 2 to 1 of the thread pitch) depending on a relative thickness of the coupling body and the tubulars and depending on the radius of the curved transitions.9. The coupling of any one of clauses 1 to 8, wherein at least a second portion of the bore (72) defines a second continuous internal thread (75b) mirroring the first continuous internal thread (75a).10. The coupling of clause 9, wherein: the first and second continuous internal threads (75a-b) meet at a center of the bore (72), whereby faces (66) on pins (64a-b) of the tubulars (60a-b) are configured to maximize pin nose bearing faces that abut one another to maximize connection torque resistance; or the bore (72) defines a reinforced area (80) disposed between the first and second continuous internal threads (75a-b), the reinforced area (80) having shoulders (84) configured to abut faces (66) on pins (64a-b) configured to maximize pin nose bearing faces (for maximum torque resistance) of tubulars (60a-b) joined by the coupling.11. The coupling of any one of clauses 1 to 10, wherein one end (71 b) of the body (70) comprises a wear sleeve (90) extending therefrom, the wear sleeve (90) having an inner wall diameter (92) and an outer wall diameter (94); and wherein a first length of the inner wall diameter (92) of the wear sleeve (90) is cylindrical, and a second length of the inner wall diameter (92) tapers outwardly in a direction away from the bore (72) of the body (70).12. A tubular system, comprising: a plurality of tubulars (60), each having pins (64) disposed on ends of the tubular (60), the pins (64) having external thread (65); a plurality of couplings (70) according to any one of clauses 1 to 11 and being configured to join the tubulars (60) together.13. The tubular system of clause 12, wherein an annular clearance limit for the tubular system downhole constrains the second outer diameter of the coupling; and wherein to the meet the annular clearance limit, each of the pins of the tubulars has a turndown recessed from the first outer diameter to a first intermediate outer diameter smaller than the first outer diameter, each of the pins tapering inward from the first intermediate outer diameter at the turndown to a second intermediate outer diameter at a pin nose of the pin, each of the pins having the external thread formed between the turndown and the pin nose, whereby both the first and second outer diameters meet the annular clearance limit.14. The tubular system of clause 13, wherein:the pipe body between the first outer diameter and the inner diameter defines: (i) a pipe tension resistance characterizing strength of the pipe body resisting yielding under tensile loading, and (ii) a pipe pressure resistance characterizing strength of the pipe body resisting yielding under pressure loading; and the connection defines: (i) a connection tension resistance characterizing strength of the connection resisting yielding under tensile loading and (ii) a connection pressure resistance characterizing strength of the connection resisting yielding under pressure loading.15. The tubular system of clause 14, wherein: the tubulars define first critical cross-sections between the first outer diameter of the pipe body and the inner diameter of the pipe body, between the first intermediate outer diameter of the turndown and the inner diameter of the pipe body, and the second intermediate outer diameter of the pin and the inner diameter of the pipe body; and the couplings define second critical cross-sections between the second outer diameter and the internal diameter of the coupling.16. The tubular system of clause 15, wherein the first and second critical cross-sections are configured to maximize at least one of: (i) the connection tension resistance of the connection relative to the pipe tension resistance of the pipe body, and (ii) the connection pressure resistance of the connection relative to the pipe pressure resistance of the pipe body.17. The tubular system of clause 15 or 16, wherein the connection tension resistance is at least 70 percent of the pipe tension resistance; and wherein the connection pressure resistance is at least 70 percent of the pipe pressure resistance.18. The tubular system of clause 15, 16 or 17, wherein the first and second critical cross-sections are selected within the annular clearance limit to maximize at least one of: (i) the connection tension resistance relative to the pipe tension resistance, and (ii) the connection pressure resistance relative to the pipe pressure resistance.19. The tubular system of any one of clause 15 to 18, wherein the pin noses of the pins are configured to meet at the center of the coupling, the pin noses having a nose bearing area configured to engage with a torque resistance characterizing engagement of the nose bearing areas resisting torque force.20. The tubular system of clause 19, wherein the first and second critical cross-sections are configured to maximize the nose bearing area of the pin noses.21 . A method of manufacturing a tubular system for use downhole, the method comprising: fabricating tubulars having pins at pipe ends; fabricating couplings having first and second ends and defining a bore therethrough; and forming a first continuous internal thread with an internal diameter in at least a first portion of the bore toward the first end of the couplings by: converging the internal diameter, for a first section of the first continuous internal thread disposed toward the first end, linearly inward toward the bore at a first angle from a first point to a first intersection; converging the internal diameter, for a second section of the first continuous internal thread connected with the first section at the first intersection, linearly inward toward the bore at a second angle from the first intersection to a second intersection, the second angle being less than the first angle; diverging the internal diameter, for a third section of the first continuous internal thread connected with the second section at the second intersection, linearly outward from the bore at a third angle from the second intersection to a second point; and defining the internal diameter of the first continuous internal thread with at least one of: (i) a first curvature transitioning between the first and second angles at the first intersection of the first and second sections, and (ii) a second curvature transitioning between the second and third angles at the second intersection of the second and third sections.22. The method of clause 21 , wherein fabricating the tubulars and the couplings further comprises: selecting a first outside diameter and a weight of a pipe body for the tubulars; selecting a second outside diameter of the couplings to fit within a predetermined annular clearance downhole; and configuring a connection at the first and second ends of the coupling to the pipe ends of the tubulars to meet a rating for tension and / or pressure by:(i) forming turndowns in the first outside diameter of the pipe body at the pipe ends of the tubulars, the turndowns recessed to a third outer diameter smaller than the first outer diameter;(ii) configuring pin thread of the pins on the pipe ends to thread respectively to the first and second continuous internal thread in the first and second ends of the coupling;(iii) configuring nose bearing area of pin noses on the pipe ends to engage near a center of the bore of the coupling; and(iv) balancing first critical cross-sections of the tubular to first critical crosssections of the coupling to produce matched-strength performance on the tubular and coupling.

[0178] The foregoing description of preferred and other embodiments is not intended to limit or restrict the scope or applicability of the inventive concepts conceived of by the Applicants. It will be appreciated with the benefit of the present disclosure that features described above in accordance with any embodiment or aspect of the disclosed subject matter can be utilized, either alone or in combination, with any other described feature, in any other embodiment or aspect of the disclosed subject matter.

[0179] In exchange for disclosing the inventive concepts contained herein, the Applicants desire all patent rights afforded by the appended claims. Therefore, it is intended that the appended claims include all modifications and alterations to the full extent that they come within the scope of the following claims or the equivalents thereof.

Claims

CLAIMS:1 . A coupling for joining tubulars, the coupling comprising: a body having a first end and a second end and defining a bore therethrough; at least a first portion of the bore defining a first continuous internal thread, the first continuous internal thread having a first section, a second section, and a third section, the first section disposed toward the first end, the second section connected with the first section at a first intersection, the third section connected with the second section at a second intersection; and an internal diameter of the first continuous internal thread for the first section converging linearly inward toward the bore at a first angle from a first point to the first intersection, the internal diameter for the second section converging linearly inward toward the bore at a second angle from the first intersection to the second intersection, the second angle being less than the first angle, the internal diameter for the third section diverging linearly outward from the bore at a third angle from the second intersection to a second point, wherein the internal diameter of the first continuous internal thread defines at least one of:(i) a first curvature transitioning between the first and second angles at the first intersection of the first and second sections, and(ii) a second curvature transitioning between the second and third angles at the second intersection of the second and third sections.

2. The coupling of claim 1 , wherein the respective at least one of:(i) the first curvature is tangential to the first and second angles; and(ii) the second curvature is tangential to the second and third angles.

3. The coupling of claim 1 or 2, wherein: the internal diameter defines the first curvature at the first intersection; the internal diameter for the first section converges linearly inward toward the bore at the first angle from the first point to the first curvature; the internal diameter for the second section converges linearly inward toward the bore at the second angle from the first curvature to the second intersection; andthe first curvature is tangential to the first and second angles.

4. The coupling of claim 1 , 2 or 3, wherein: the internal diameter defines the second curvature at the second intersection; the internal diameter for the second section converges linearly inward toward the bore at the second angle from the first curvature to the second curvature; the internal diameter for the third section diverges linearly outward from the bore at the third angle from the second curvature to the second point; and the second curvature is tangential to the second and third angles.

5. The coupling of claim 4, wherein the first curvature is tangential to the first and second angles; and wherein the second curvature is tangential to the second and third angles.

6. The coupling of any one of claims 1 to 5, wherein the first continuous internal thread is configured to thread to a pin of one of the tubulars; and wherein: the pin having an external buttress-type thread; or the pin having a pin thread with a single taper angle, and at least one of the first angle, the second angle, and the third angle of the first continuous internal thread is approximate to the single taper angle of the pin thread, optionally wherein the second angle of the second section is approximate to the single taper angle of the pin thread on the pin of the tubular.

7. The coupling of claim 1 , wherein the at least one of the first and second curvatures spans a helical profile around the first continuous internal thread where one of the sections transitions to the adjacent section.

8. The coupling of claim 7, wherein: the helical profile, of the first continuous internal thread being formed by a threading tool, spans between a half (T ) rotation to one (1 ) full rotation of the coupling as the coupling is threaded by the threading tool; the first continuous internal thread has a thread pitch, and an axial length of the helical profile is between 1 / 4 to 1 / 2 of the thread pitch to 1 / 2 to 1 of the thread pitch; or an extent of the helical profile is defined by (i) a relative thickness of the body and the tubulars and (ii) radii of the first and second curvatures.

9. The coupling of any one of claims 1 to 8, wherein at least a second portion of the bore defines a second continuous internal thread mirroring the first continuous internal thread.

10. The coupling of claim 9, wherein: the first and second continuous internal threads meet at a center of the bore, whereby faces on pins of the tubulars are configured to maximize connection torque resistance between the faces that abut one another; or the bore defines a reinforced area disposed between the first and second continuous internal threads, the reinforced area having shoulders configured to abut faces on pins of the tubulars joined by the coupling, the faces abutting the shoulders being configured to maximize torque resistance.11 . The coupling of any one of claims 1 to 10, wherein one end of the body comprises a wear sleeve extending therefrom, the wear sleeve having an inner wall diameter and an outer wall diameter; and wherein a first length of the inner wall diameter of the wear sleeve is cylindrical, and a second length of the inner wall diameter tapers outwardly in a direction away from the bore of the body.

12. A tubular system, comprising: a plurality of tubulars, each of the tubulars having a pipe body with a first outer diameter and defining a first bore with an inner diameter, each of the tubulars having pins disposed on ends of the tubular, the pins having external thread; and a plurality of couplings being configured to join the tubulars together, each of the couplings comprising a coupling body having opposing ends, having a second outer diameter, and defining a second bore therethrough, the second outer diameter of the couplings being greater than the first outer diameter of the pipe body, the second bore defining continuous internal threads extending outward from a center of the second bore respectively to the opposing ends of the coupling body, each of the continuous internal threads of the couplings being configured to thread in a connection to a respective one of the external threads of the pins, each of the continuous internal threads having a first section, a second section, and a third section, the first section disposed toward one of the opposing ends, the second section connected with the firstsection at a first intersection, the third section connected with the second section at a second intersection, an internal diameter of each of the continuous internal threads for the first section converging linearly inward toward the second bore at a first angle from a first point to the first intersection, the internal diameter for the second section converging linearly inward toward the second bore at a second angle from the first intersection to the second intersection, the second angle being less than the first angle, the internal diameter for the third section diverging linearly outward from the second bore at a third angle from the second intersection to a second point, wherein the internal diameter of each of the continuous internal threads defines at least one of:(i) a first curvature transitioning between the first and second angles at the first intersection of the first and second sections, and(ii) a second curvature transitioning between the second and third angles at the second intersection of the second and third sections.

13. The tubular system of claim 12, wherein each of the couplings is configured according any one of claims 2 to 11 .

14. The tubular system of claims 12 or 13, wherein an annular clearance limit for the tubular system downhole constrains the second outer diameter of the couplings; and wherein to meet the annular clearance limit, each of the pins of the tubulars has a turndown recessed from the first outer diameter to a first intermediate outer diameter smaller than the first outer diameter, each of the pins tapering inward from the first intermediate outer diameter at the turndown to a second intermediate outer diameter at a pin nose of the pin, each of the pins having the external thread formed between the turndown and the pin nose, whereby both the first and second outer diameters meet the annular clearance limit.

15. The tubular system of claim 14, wherein: the pipe body between the first outer diameter and the inner diameter defines: (i) a pipe tension resistance characterizing strength of the pipe body resisting yielding under tensile loading, and (ii) a pipe pressure resistancecharacterizing strength of the pipe body resisting yielding under pressure loading; and the connection defines: (i) a connection tension resistance characterizing strength of the connection resisting yielding under tensile loading and (ii) a connection pressure resistance characterizing strength of the connection resisting yielding under pressure loading.

16. The tubular system of claim 15, wherein: the tubulars define first critical cross-sections between the first outer diameter of the pipe body and the inner diameter of the pipe body, between the first intermediate outer diameter of the turndown and the inner diameter of the pipe body, and the second intermediate outer diameter of the pin and the inner diameter of the pipe body; and the couplings define second critical cross-sections between the second outer diameter and the internal diameter of the coupling.

17. The tubular system of claim 16, wherein the first and second critical crosssections are configured to maximize at least one of: (i) the connection tension resistance of the connection relative to the pipe tension resistance of the pipe body, and (ii) the connection pressure resistance of the connection relative to the pipe pressure resistance of the pipe body.

18. The tubular system of claim 16 or 17, wherein the connection tension resistance is at least 70 percent of the pipe tension resistance; and wherein the connection pressure resistance is at least 70 percent of the pipe pressure resistance.

19. The tubular system of claim 16, 17 or 18, wherein the first and second critical cross-sections are selected within the annular clearance limit to maximize at least one of: (i) the connection tension resistance relative to the pipe tension resistance, and (ii) the connection pressure resistance relative to the pipe pressure resistance.

20. The tubular system of any one of claims 16 to 19, wherein the pin noses of the pins are configured to meet at the center of the coupling, the pin noses having a nose bearing area configured to engage with a torque resistance characterizing engagement of the nose bearing areas resisting torque force.21 . The tubular system of claim 20, wherein the first and second critical crosssections are configured to maximize the nose bearing area of the pin noses.

22. A method of manufacturing a tubular system for use downhole, the method comprising: fabricating tubulars having pins at pipe ends; fabricating couplings having first and second ends and defining a bore therethrough; and forming a first continuous internal thread with an internal diameter in at least a first portion of the bore toward the first end of the couplings by: converging the internal diameter, for a first section of the first continuous internal thread disposed toward the first end, linearly inward toward the bore at a first angle from a first point to a first intersection; converging the internal diameter, for a second section of the first continuous internal thread connected with the first section at the first intersection, linearly inward toward the bore at a second angle from the first intersection to a second intersection, the second angle being less than the first angle; diverging the internal diameter, for a third section of the first continuous internal thread connected with the second section at the second intersection, linearly outward from the bore at a third angle from the second intersection to a second point; and defining the internal diameter of the first continuous internal thread with at least one of: (i) a first curvature transitioning between the first and second angles at the first intersection of the first and second sections, and (ii) a second curvature transitioning between the second and third angles at the second intersection of the second and third sections.

23. The method of claim 22, wherein defining the internal diameter comprises defining the internal diameter with at least one of:(i) the first curvature tangential to the first and second angles; and(ii) the second curvature tangential to the second and third angles.

24. The method of claim 22 or 23, wherein fabricating the tubulars having the pins at the pipe ends comprises: forming an external buttress-type thread on the pins, the first continuous internal thread being configured to thread to the external buttress-type threads on the pins; orforming pin thread with a single taper angle on the pins, and forming the first continuous internal thread by forming at least one of the first angle, the second angle, and the third angle of the first continuous internal thread approximate to the single taper angle of the pin thread.

25. The method of claim 22, 23 or 24, wherein defining the internal diameter with the at least one of the first and second curvatures comprises spanning the at least one of the first and second curvatures along a helical profile of the first continuous internal thread.

26. The method of claim 25, wherein spanning the at least one of the first and second curvatures along the helical profile comprises: spanning between a half (T ) rotation to one (1 ) full rotation of the coupling as the coupling is threaded by a threading tool used to form the first continuous internal thread; or defining an axial length of the helical profile between 1 / 4 to 1 / 2 of a thread pitch to 1 / 2 to 1 of the thread pitch for the first continuous internal thread; or spanning an extent of the helical profile defined by (i) a relative thickness of the coupling and the tubulars and (ii) radii of the first and second curvatures.

27. The method of any one of claims 22 to 26, further comprises forming a second continuous internal thread with an internal diameter in a second portion of the bore toward the second end of the couplings such that the second continuous internal thread mirrors the first continuous internal thread.

28. The method of claim 27, wherein forming the first and second continuous internal threads comprises: meeting the first and second continuous internal threads at a center of the bore, whereby faces on pins of the tubulars are configured to abut one another; or defining a reinforced area between the first and second continuous internal threads, the reinforced area having shoulders configured to abut faces on pins of the tubulars joined by the coupling.

29. The method of any one of claims 22 to 28, wherein fabricating the tubulars and the couplings further comprises: selecting a first outside diameter and a weight of a pipe body for the tubulars; selecting a second outside diameter of the couplings to fit within a predetermined annular clearance downhole; andconfiguring a connection at the first and second ends of the couplings to the pipe ends of the tubulars to meet a rating for tension and / or pressure by:(i) forming turndowns in the first outside diameter of the pipe body at the pipe ends of the tubulars, the turndowns recessed to a third outer diameter smaller than the first outside diameter;(ii) configuring pin thread of the pins on the pipe ends to thread respectively to the first and second continuous internal thread in the first and second ends of the couplings;(iii) configuring nose bearing area of pin noses on the pipe ends to engage near a center of the bore of the couplings; and(iv) balancing first critical cross-sections of the tubulars to first critical cross-sections of the couplings to produce matched-strength performance on the tubular and coupling.

30. The method of claim 29, wherein configuring the pin thread to thread to the first and second continuous internal thread and configuring the nose bearing areas to engage comprises: forming the pin thread externally on the pipe ends tapering inward from the third outer diameter at the turndowns to a fourth outer diameter at the pin noses of the pipe ends, and forming first and second sections of the first and second continuous internal thread internally on the bore of the coupling and tapering from a center of the bore to opposing ends of the coupling.31 . The method of claim 29 or 30, wherein configuring the connection by balancing the critical cross-sections to produce matched-strength performance on the tubular and coupling further comprises reducing hoop stresses and displacements of the pins of the tubulars and the ends of the couplings.

32. The method of claim 29, 30 or 31 , wherein configuring the connection by balancing the critical cross-sections comprises configuring the connection to meet the rating for both tension and pressure by balancing the critical cross-sections of the pipe body, the turndown, and the pin to the critical cross-sections of the coupling, the balanced critical cross-sections being configured to achieve targeted connection tension and pressure efficiencies relative to the pipe body.

33. The method of claim 32, wherein configuring the connection to meet the rating for both tension and pressure comprises configuring the rating for both tension andpressure of the connection to be at least 70 percent of a minimum tension and pressure rating of the pipe body of the tubulars.

34. The method of any one of claims 29 to 31 , wherein configuring the connection by balancing the critical cross-sections comprises configuring the connection to meet the rating for pressure alone by balancing a critical pressure cross-section of the coupling to be as large as possible relative to a critical cross-section of the pipe body within constraints of: (i) a wall thickness of the pipe body and (ii) the second outside diameter of the coupling.

35. The method of any one of claims 29 to 31 , wherein configuring the connection by balancing the critical cross-sections comprises configuring the connection to meet the rating for tension alone by maximizing a ratio of a critical tension cross-section of the coupling relative to a critical cross-section of the pipe body within constraints of: (i) a wall thickness of the pipe body and (ii) the second outside diameter of the coupling.

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