WO2013033684A1

WO2013033684A1 - Drill-pipe tool-joint

Info

Publication number: WO2013033684A1
Application number: PCT/US2012/053589
Authority: WO
Inventors: John Watts
Original assignee: WATTS RAMOS, Beverly
Priority date: 2011-09-04
Filing date: 2012-09-03
Publication date: 2013-03-07

Abstract

Features for drill-pipe tool-joints are disclosed that extend service life and reduce the OD for use in drilling of shale formations.

Description

DRILL-PIPE TOOL- JOINT

US 6378,880, Application PCT/US2007/001 154 and paper SPE f ADC 29352 are made a part of this application by reference. Priority is claimed on Provisional Applications 61/530,981 and 61/540,359

FIELD OF THE INVENTION : Drilling of Oil, Gas and Water Weils. BACKGROUN D ART

Definitions for purposes of this application are as follows: Box ^~ An internal pipe thread formed in a first end of a joint of a pipe; Pi ^:::: An external pipe thread formed on a second end of a joint of pipe; Critical stress ~ The least cyclic stress in a given material that can cause fatigue failure, usually considered to he one-half of that materials yield stress; A thread load - the load per tangential unit length transmitted by a given thread to its mating thread; A critical thread load ^:::: The least thread load transmitted by a given thread to its mating thread that can cause a critical stress at the root of its mating thread; Axial stress ^::: The ax ial load on an assembled tool-joint at a given point along the axis intermediate the largest diameter thread and the smallest diameter thread, divided by its cross-section wall area at that point; Root stress ^:::: the radial bending stress at a thread root multiplied by the stress concentration factor (SCF) at that root; Maximum shear stress ⁵⁵⁵ (The axial stress, minus the radial root stress) /2. where. Tension is plus and.

Compression is minus; OD ^::: outermost diameter; ID ^::: innermost diameter; RSC ^:::: rotary shouldered connection; Flank angle - the angle between a flank and a plane perpendicular to the axis. Tool-Joint ^:::: a box and pin that connects joints of drill-pipe pipe to form drill strings for drilling wells into the earth; Trapped thread = a thread having a negative included angle between flanks, the axial width between flanks being greater at the root than at the crest; Open thread - a thread having an included angle not less tha zero between flanks, the axial gap width between flanks being equal or less at the root than at the crest. A wedgethread has a load flank pitch greater than its stab flank pitch, the effective center pitch being equal to the average of the flank pitches.

A conventional tool-joint is a box and pin on ends on joints of drill pipe as shown in FIG. I, having tapered pin threads (I) that mate with tapered box threads (2), both pin and box having face shoulders (3,4) at the large diameter ends of the threads, the shoulders being dimensioned to butt against each other during assembly, causing friction between, the shoulders and also between the thread load flanks that resist the increasing makeup torque which causes the pin neck (5), between the shoulder and the engaged threads, to stretch axially while the adjacent box wail (6) compresses axially until the desired input torque is reached. Thus, the pin neck is preloaded axially in tension and the box shoulder and load flanks are preloaded axially in compression, solely by input torque, to hopefully be less subject to fatigue failure than were it not preloaded, because the stress cycle amplitude is reduced by the preload. However, the threads do not seal, so the magnitude of the make-up torque must be high enough to make and maintain a fluid seal between shoulders, and also be higher than any service torque so as to not loose in service due to combinations of mechanical or thermal service loads. Also the stress cycle amplitude at the fillet of the pin shoulder (7) and the fillet at the root (8 ) of largest diameter pi n thread is intended to be sufficiently low to prevent fatigue failure. Most tool-joints are now friction welded to the drill-pipe with a box on one end and a pin on the other end so the tool-joint walls can taper thicker from the drill pipe, to the necks of the box and pin. Also the box wall tapers thinner with the thread angle toward box face shoulder (4) which must be wide enough to withstand the high compression force between the box and pin shoulders during makeup, irrespective of other lim tations on box wall thickness.

Thread forms of conventional tool-joints have high flank angles such as 30 degrees., that expand the box outwardly and compress the pin inwardly as assembly and service loads are applied, such that the pin. and box wails must be formed much thicker than the drill-pipe walls to hopefully absorb shoulder pre-loads and various combinations of sendee loads without flexing. Conventional tool-joint OD's (^"9, 10) are made much greater than the drill- pipe OD's (1 1 ,12), which increases bending stresses in the drill-pipe adjacent the tool joint as at the juncture ( 18) because the long and large OD of the tool-joint does not absorb its proportional share of the bending exerted on the string, and .it also acts as a lever to bend the drill-pipe at the juncture back and forth with each revolution, which accelera tes fatigues failure of the drill-pipe adjacent the tool-joint . as influenced by the environment and the cyclic stress magnitude,

A conventional tool joint OD is made very large to resist fatigue failure, but it worsens other problems including: (t) increased OD wear as the large OD holds the drill pipe away from the bore hole wall, taki ng wear that, would otherwise be distributed along 30 feet of pipe; (2) key- seating, which is a terra for the tool-joint becoming grooved and stock in a side of the hole wall that often prevents raising or lowering the pipe; (3) preventing the drill string from, entering smaller diameters down-hole; (4) increasing resistance to flow of the drilling mud as it returns up a narrow amiuius betwee the tool-joint OD and the ID of the hole it is within (5) increasing drill-pipe bending stresses which causes fatigue failures in the drill pipe where it is connected to the tool-joint; and (6) it greatly increases bending stresses in the drill pipe as it passes through ball joints at the top and bottom of risers used in offshore drilling.

As a conventional tool-joint rotates in the well bore, its larger OD rubs the hole wali which accelerates wear on the tool joint OD due to the higher peripheral speed and also on the ID of the casing it is within, especially when in a dogleg, which necessitates a hard band (13) around it such as carbide coating to resist such wear. Pipe tongs cannot grasp the hard hand, so the tool joint must be made longe than otherwise necessary to provide long space for tightening the tool-joint which further increases bending stress adjacent the tool joint and in turn, the longer tool joint acts as a lever to further increase bending stress on the tool-joint and on the drill pipe when in a dogleg. St has long been known thai the first engaged thread on the tension side of a screw thread is stressed much higher than the other threads, be it a bolt thread or pipe thread, and that the increased load tends to cause fatigue failure at that point. Additionally , stretching of the pin neck and compressing of the adjacent box as described above, often exerts an extreme thread bending stress higher than the critical stress, at the root of the largest diameter engaged pin thread (14) that can combine with stress fluctuations to cause pin neck fatigue failure, as described ne t.

Some excerpts from an extensive study on drill-pipe failures, Publication SPEIADC 29352 are; "RSC optimum makeup torques are different under static and dynamic (fatigue) loading conditions— friction capacity depends on different factors and may vary within a large range (more than. 50%) during connectio life and from one assembly to another— friction capacity of used copper plated connections under temperature (260 deg C) was three times as much as for the new one— An effective operative method for the evaluation of the optimum makeu torque as a function of the previously mentioned factors does not exist— if RSC friction capacity is unknown, insufficien makeup torque may result in: RSC separation during running, washout (between the shoulders), fatigue failure, stretched pins, bellied boxes or even twist-off while drilling... Failure analysis showed that the main problems were the following; RSC fatigue failure (53%); washout through RSC shoulder faces (38%); twist-off (9%)."

Thus it is very clear that basic problems have not been, solved during the last hundred years and additionally, there has recently arisen needs for tool joints having even much higher torque ratings, smaller OD's and larger ID's than conventional tool joints, for use in the drilling of horizontal gas wells through shale formations that exist over large areas of the world, to produce vast quantities of clean burning gas said to be a 200 year supply. However higher torques, combined with smaller OD's and larger ID's, are contradictory to the conventional tool- joint design methodology described above, so new invention is now required. Fatigue failure in a conventional tool-joint typically occurs near the largest diameter engaged pin thread (14) and at the juncture (I S) of the tool-joint and the dill pipe when the cyclical stress there, exceeds the critical stress due to excessive length (17) of the OD of the tool-joint. Fatigue failure usually occurs at the largest diameter engaged pin thread (14) because the stress differential across that thread varies as cited in the SPE paper since stress is greater than across any other thread and because the box wall there (6), is thick enough to transmit a critical load as taught herein; Fatigue failures also occur in the box wall neck (15) at the box smallest engaged thread ( 16) but rarely, because the bending stress there is less than at the pin neck, being on a smaller diameter. The largest pin thread, of conventional tool-joints is pre-loaded higher than its other threads because of the neck stretch, and because the axial length between threads are not dimensioned to load the threads evenly, and the box. wall at that point is thick enough to transfer more than a critical load. Although the amplitude of the cyclical stress is reduced by preloading the shoulders, the mean stress is always increased into the critical stress range.

Premium tubing connections having old type "trapped wedge threads" have been used for brief light duty drilling to gain the torque advantage of a wedge thread, but they have not been recommended for regular or heavy-duty drilling operations due to their weakness and short service lives when used as tool-joints, having been designed for static stresses. They have negative included angles, so the thread dope that is supposed to flow out from between mating threads during makeup, can be trapped between roots and crests and pressurized, which wedges the negative flanks against each other radially and stops makeup before the position, of full makeu is attained, allowing the connection to loosen as the dope further oozes out over a period of time, and a. loose tool joint typically results in a dropped drill string. Further, such tubing connections have thin walls that have extremely short wear life as the connection rubs against the borehole wall The negative stab flank angles make stabbing, and makeup of the pin. into the box, slow and tedious, which increases running costs and the probabi lity of failure. As assembled, the pin thread is longer than the bo thread, which creates extreme bending stresses against stab flanks near the pin face, and also cause extreme bending stresses on load flanks near the box face, both causing criticaJ stresses that lead to fatigue failure. Such stresses as assembled., prevent the box and pin from recei ing stresse as an integral membe but instead, for instance, when a service tension load is applied, axial tensio stress is increased in the pin and

compressive stress is reduced in the box, which precludes an orderl stress distribution necessary to prevent fatigue failure.

DISCLOSURE OF THE INVENTION

To prevent fatigue failure in a Tool Joint of the present invention, the cyclic stress at even' thread is made less than the critical stress. The axial stress may be reduced by increasing the box and pin neck areas and also the combined cross-sectional area of the box and pin threads to be greater than the cross-sectional area of the pipe. The root stress may be reduced by:

decreasing the thread form depth-io-width-raiio; dimensioning the box arid pin threads to have radial interference upon assembly; increasing the helical thread length; by forming the pin threads shorter in axial length than the box threads in accord with Poisson's Ratio, as taught in my patent OS 12/087,762, so as to be of equal length after assembly together and thereby react to service loads as though the box and pin were integral. The length of the tool joint QD should be only Song enough for the box, the tong surface lengths may be on a smaller diameter to be more flexible and to reduce bending stresses on the threads and junctures. The tool-joint should be dimensioned so the axial stress, the thread root bending stress and combined stresses are all less than the critical stress.

The present invention also teaches: (1) Dimensioning box and pin threads to have radial interference, with pin threads having shorter leads than box threads to counteract lengthening of the pin thread lead and shortening of the box thread lead thai occurs upon assembl due to

Poisson's Ratio, to thereby provide an assembled box and pin having equal thread leads not stressed axia!ly against each, other, which reduces stresses- on all threads such that the box and pin react to service loads as if they were of one piece; (2) Dimensioning the thread form to have a smaller depth-to- width ratio to further reduce bending stress at the threat! root; (3) Use small flank angles that eliminate slippage between mating threads and lower stresses in the bo and pi walls; (4) Make neck walls of both box and pin thicker than the drill pipe wall to reduce the axial stress sufficiently to prevent, critical stresses when that axial stress combines with the root bending stress; (5) Dimension box and pin wall thicknesses at their faces thin enough to prevent threads from transmitting critical loads to their mating threads, (Load equals stress multiplied by area, so the thinner the wall near faces, the lower the thread load.) (6) Reduce the maximum stress at the roots so that when the bending stress combines with the axial stress, the combined stress will be below the critical stress; (7) Form stress relief grooves adjacent the end engaged threads to reduce the stress concentration factor at roots of the end threads. Thus when T-Tenston Stress and C - compression stress, the elements of a tool-joint under tension have maximum shear stresses at the root of the stab flank ^:::: (T-(~C)}/2 ^::: (T-s-C)/2₅ and the elements of a tool-joint under compression have a maximum shear stress at the root of the stab flank ~ (-C- (T))/2 ~ -(C+T)/2, so the reversal of stress at the root of the stab flank goes from (T+C }/2 to - (C+T)/2. Low ratios of thread depth to thread width, tend to prevent fatigue failure by reducing bending stress.

Open type wedge threads taught: by my Patent '880 provide the high torque resistance and also the fluid seal required without need for large diameter butting shoulders of conventional tool-joints or pin-nose seals, and they also avoid failures caused by variations in the coefficient of friction as explained i n the SPE paper above. This invention also teaches that the pin thread lead is formed shorter than the box thread lead before assembly, per Ί 154, so they will be equal upon their assembly together, unlike conventional tool joints whose box and pin pitch are equal as machined, but unequal after assembly, leaving the thread highly stressed in bending even before service loads are applied. The threads taught by Ί 154 wedge tightly together upon assembly which provides high torque and also prevents loosening between mating threads that causes fretting and accelerated fatigue found in conventional threads having high flank angles. Of major importance is the fact that my wedge type threads become progressively wider from face to neck, such that threads at the pi and box necks, where fatigue failure tends to occur, are so much wider that high critical bending stresses are not generated at the largest diameter pin thread or the smallest diameter box thread whereas the mating thread is formed on a thin wail and transmits a small load which together, greatly reduce root bending stress. Since my box and pin thread leads are equal upon assembly, they exert no bending stresses against each other as assembled before service loads are applied, so the pi and box receive stress together as one piece of steel when torqued, stretched or compressed. Thus, the axial service load that each thread can transmit to its mating thread is limited by the least radially thick cross-sectional area of the box or pin at that point, such that a connection can be dimensioned with none of its threads being able to transmit a critical load.

It is also a feature of this invention to substantially dimension the tool joint OD no larger, and the ID no smaller than is necessary to include the other features required for a given service and to that end, a thread taper may be selected to provide wall tapers required to keep thread loads below critical loads for each and every thread, such that box OD and pin ID may be formed cylittdricaily as shown in FIG 3, to reduce the box OD and increase the box and pin ID, which in turn, reduces costs to make and use it, The pin neck and box neck should be made thick enough to reduce axial, stresses sufficiently to preclude critical stresses when combined with any root stress, but the thinnest of the two walls at any given point must not be thick enough to transmit a critical thread load t the other wall; the number of engaged threads should be sufficient to withstand the desired torque rating and axial load rating of the tool joint; the box and pin walls should taper down slowly enough to not cause stress risers; tong grip space may be provided on a diameter intermediate the box OD and the pipe OD; a stress relief groove may be formed adjacent the largest diameter pin thread and one adjacent the smallest box thread, and box and pin bores should taper gradually to connect with the dill pipe without forming stres risers. The absence of butting shoulders precludes the overstressing of the first engaged thread upon assembly, as occurs with conventional tool-joints.

Pipe ends may be upset to attain the thicker walls so tool-joints may be friction welded on ends of drill-pipe joints in accord with current practice. The Tool-Joint may be dimensioned at the end to provide a full bore connection that can be a great advantage to use in such as casing drilling operations. Features of the invention may be used in any combination as may be required for specific services.

My provisional Patent Application '35 discloses the use of hard bands of metal applied around a tool-joint OD to reduce wear as it rubs against the bore hole waif the tool-joint threads having little of no radial interference between the box and pin thread such that the hard band material is not caused to expand, crack and crumble when the tool-joint is assembled.

A second hard band configuration is to apply the hard material in short segments around, the tool-joint (3D, segments being axiall displaced from on another hut overlapped

circumferential!}', the segments being too short to crack, which may be applied around a box that does expand when assembled with a pin.

Such hard band configurations may be applied around a tool-joint OD at any point along its length, even near the box face where wear is most critical because it is the thinnest portion. BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a quarter section of a conventional tool joint box and pin assembled together.

FIG, 2 is a quarter section depicting a. box and pin of the present invention.

FIG. 3 ts a quarter section depicting tbe present invention having cyiindrical pin ID and box OD.

FIG. 4 is a detail of the thread form of the present invention.

FIG. 5 is a quarter section of the present invention that depicts additional features.

FIG. 6 depicts a hard band configuration.

PREFERRED EMBODIMENTS OF THE INVENTION

A preferred embodiment shown in FIG 2 comprises a pin (22) having tapered pin threads (24), a pin inner tapered wall (26), a pin face (28), largest diameter engaged pin thread (20), smallest diameter engaged pin thread (27) a pin neck OD (21 ), a pin neck ID (23), and a pin neck (25) between the pin OD and ID. A Mating box (29) has box threads (34), a tapered box outer wall (36), a box face (38), a smallest box thread (37), a box neck OD (30), a box neck ID (39), and a bo neck (32) between the box neck OD and ID. The thread form in FIG. 4, may have positive stab flank angle (43), negative load flank angle (42) and zero degree included angle (44). The box wail thickness of FIG. 2 increases from the box face toward the box neck and the in wall thickness increases from the pin face toward the pin neck, neither wall, being thick enough at any given point up to its mid-length to transmit a critical thread load to the adjacent wail. The pin and box threads are dimensioned to interfere radially sufficiently to seal between roots and crests, and the pin thread axial pitch (3.1) is formed shorter than the box thread axial pitch (33) in accord with Poisson's Ratio, so the leads become substantially equal upon assembly, and react to service loads and temperatures, as though the assembled box and pi were one piece without substantial, bending stresses on the threads. The thinner wall of the box and pin at any axial point, is dimensioned too thin to transmit a critical thread load. When using ray wedge type thread, the thread width (40) but not the thread depth (41) in FIG 4, increases progressively from the face toward the neck which reduce the bending stress at the stab root (45) and the load root (46) as the thread load per tangential inch, increases toward the axial point where the two walls are of equal thickness. Thus, the walls of both the box and pin increase in thickness from their face toward their mid-length but are not thick enough to transmit a critical load to any mating thread. I the direction from thread mid-length toward either face, both the loads and the wall cross-section area is reduced as each thread on the thinner wall transmits its load to its mating thread, so the resulting axial stress at each thread, and the bending stress at each thread must be determined and combined, to insure that the highest stress there, is less than the critical stress, A stress analysis should be performed on each thread of both the pin and the box. Design Factors such as: Rated drill pipe axial loads in compression, tension, bending and torque; desired, tool joint OD and ID; and the critical stress for the material must be determined. Before designing a tool-joint a sufficient thread form, the tool -join t material and ratings should be selected and the

to Stress Concentration Factor (SCF) should be determined prior to dimensioning the wall thicknesses.

After required wai! thickness and tapers of the box and pin are determined, the thread taper may be adjusted to make the tool joint OD (70) and ID (71 ) cylindrical per FIG 3, to maximize the ID and to minimize the OD. Between the box neck and its attachment to the drill pipe (73), and between the pin neck and its attachment to the drill pipe (74k tapers (75,76) may be formed to extend for attachment to ihe drill pipe by any suitable means, and the pi may be formed with ID (77) and the box ID may be formed with a diameter (78), both JD's substantially equal to pin ID to form a flow passage for drilfiug muds to have an acceptable pressure drop. A hard band (79) is deposited around periphery (80) of the box. The thread axiai pitch, the thread dept and/or the number of threads may be adjusted, to reduce the bending stress if necessary. Although this invention is intended initially for use o drill pipe, it may also be used in other connection services such as casing drilling, bolting and industrial power shafts. I have found that a thread taper of . 14 (change of diameter per unit length) works well, that tapers less than JO result, in stabbing problems and excessive thread lengths and that tapers greater than .25 do not allow a sufficient number of threads within radial limits. As shown in FIG 5 tong surface (SI.) may be formed on the pin, and tong surface (52) is formed on the box, both tong surfaces having diameter intermediate the box OD (53) and the pipe OD (54). The tong surface diameters should be enough larger than the pipe OD to accept tong wear and still be stronger than the pipe.

Should conventional hard bands, be required as at (55) in FIG. 5, then the box OD length (64) between the thread (56) of the box threads and taper (57), may be extended. Thus, the box and pin tong surfaces are reduced in diameter which makes the tool -joint much more flexible than conventional tool-joints and substantially reduces bending stresses on the drill pipe to tool-joint juncture (58) to prevent fatigue failure there. Engaged threads may extend from the largest engaged pin thread (59) to the smallest engaged bo thread (56).

The preferred embodiment shown in FIG 5 includes ID'S (61 ,62) through the box and pin being substantially equal to the pipe ID to reduce pressure loss in the drilling mud and to allow passage of logging tools and such. Box taper (60) joins the tool-joint juncture (58) with tong surface (52) and pin taper (65) joins pipe OD (54) with tong surface (51).

As a further precaution against critical stresses, a smooth surface and large radius stress relief groove (63) may be formed in the pin neck adjacent the largest engaged pin thread (59). Alternate hard band segments (102.104,120, 121) as depicted io FIG. 6 may be applied anywhere along the OD tool-joint box (100), even near the box face because the hard bands are segmented to prevent cracking of the hard hands which permits mi nimum tool-joint length and minimum bending stress which in torn, improves service Hfe of the drill string. The segments may be overlapped as at ( 106) to provide continuous hard material around periphery (1 10) of the box.

Claims

ί claim:

1. A Tool-joint having a pin (22) a pin neck (25), a pin face (28), a largest diameter engaged pin thread (20), the connection also having a box (29) with a box neck (32 ) a box face (38), a box thread (34) formed within the box that mates with the pin thread, the pin neck and and the box neck each having a greater cross-sectional areas than the drill pipe, the box wall tapering thinner from its neck toward its face, the pin wall tapering thinner from its neck toward its face, comprising:

The box wall being formed too thin between its face and the mid-length engaged thread to transmit a critical thread load to the mating pin threads,

2. The tool-joint of claim 1 also having a smallest diameter engaged pin thread (27) that mates with the smallest box thread (37) further comprising:

The pin wall being formed too thin between its face and the mid-length engaged thread to transmit a critical thread load to the mating box threads.

3. A tool-joint having a thread form with a thread dept (41), a thread width (40) a positive stab flank angle (43) and a negative load flank angle (42), comprising:

The pin thread depth at. the largest diameter engaged thread being no more than half of the thread width; the box thread depth at the smallest diameter engaged box thread being no more than half of the thread width; neither flank angle being more than 20 degrees.

4. A tool-joint having a wedge-type thread positive stab flank angle (43), a negative load flank angle (42) and an included angle (44) formed between the stab flank and the load flank, comprising:

The included angle being not less tha zero degrees.

5. The tool-joint of claim 1 or claim 2 wherein the box and pin threads are formed to interfere radially upon assembly, the box threads having an axial pitch (33) and the pin threads having an axial pitch (31 ), further comprising: The pin thread pitch being formed shorter than the box thread pitch in accord with Poisson's Ratio such that upon assembly, the box thread pitch substantially equals the pin thread pitch.

6, The tool-joint of claim 1 or 2 further comprising: the pin thread width (40) at the largest diameter thread being at least twice the thread depth ( 3): the box thread width at the smallest diameter thread being at least twice the thread depth; neither flank angle being more than 20 degrees: an included angle (44} formed between the stab flank and the load flank; the included angle being not less than zero degrees; the effective pin thread axial pitch (31) being formed shorter than the effective box thread axial pitch (33) in accord with Poisson's Ratio such that upon assembly, the box pitch substantially equals the pin pitch,

7, A tool-joint having a box OD (53) that i larger than, the pipe OD (54), a tong OD (52) and the pin having a tong OD (51 } comprising:

The box tong OD being dimensioned intermediate the box OD and the pipe OD; the pin having a tong OD being dimensioned intermediate the box OD and the pipe OD.

8, A Tool-Joint of claim 7 having a length (64), further comprising:

The combined box and pin OD having a length less than twice the box thread length,

9, A tool joint having box threads (34) and pin threads (24) dimensioned to have radial interference between them whe assembled, the tool joint having hard band segments (102, 104,1.20.122) around its periphery (1.10), comprising:

The hard bands being segmented so as to have no segment long enough to crack upon assembly of the tool-joint.

10, A tool-joint having mating bo and pin formed with mating threads, the bo having a hard band (79) around its box periphery (80), comprising;

The threads being dimensioned to not have enough radial interference to cause the hard band to crack upon assembly of the tool-joint. 1 ί . The tool-joint of claim (1-4, 7- 10} being formed with wedge threads.

12. The tool -joint of claim ( -4, 7-10} being formed with open wedge threads, the thread groove being at least as wide at the crest as it is at the root.