LINEARLY ACTUATED, PIVOTING, PUMPING UNIT

Description

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of U.S. Provisional Patent Application No. 63/641,295, filed on May 1, 2024, U.S. Provisional Application No. 63/692,556 filed on Sep. 9, 2024, and U.S. Provisional Application No. 63/740,664 filed on Dec. 31, 2024 the entire disclosures of which are herein incorporated by reference.

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

Not applicable.

REFERENCE TO APPENDIX

Not applicable.

BACKGROUND OF THE INVENTION

The present disclosure generally relates to a linearly actuated, pivoting, beam pumping unit that for use in a downhole artificial lift system of the type that may be used to remove hydrocarbons from the ground.

It is to be understood that the discussion above is provided for illustrative purposes only and is not intended to and does not limit the scope or subject matter of the appended or ultimately issued claims or those of any related patent application or patent. Thus, none of the appended claims, ultimately issued claims or claims of any related application or patent are to be limited by the above discussion or construed to address, include, or exclude each or any of the above-cited features or disadvantages merely because such were mentioned herein.

BRIEF SUMMARY OF THE INVENTION

A brief summary of the inventions indicating their nature and substance may be understood from the subject matter presented in the claims filed with this application, which are incorporated into this brief summary by reference for all purposes, and by the inventions presented in any claims that may be issued from this application, which claims also are incorporated into this brief summary by reference for all purposes.

In one exemplary embodiment, the presently disclosed subject matter may take the form of a drive system for stroking a sucker rod, where the drive system comprises: a support base; a walking beam; a linear actuator having a lower end and an upper end, the linear actuator being adapted to expand and retract in length; an upper bearing coupling the upper end of the linear actuator to the walking beam, such that the actuator can pivot with respect to the walking beam as the linear actuator expands and retracts in length; and a lower bearing coupling the lower end of the linear actuator to the support base, such that the linear actuator can pivot with respect to the support base as the linear actuator expands and retracts in length.

Alternatively, or additionally, an embodiment of the present disclosure may take the form of a drive system for stroking a sucker rod coupled to a sucker rod pump comprising: a walking beam having a longitudinal length, a first end, and a second end; a horse head coupled to the first end of the walking beam; a support base, the support base comprising a first horizontal section extending in a direction substantially perpendicular to the longitudinal length of the walking beam and a second horizontal section extending substantially parallel to the first horizontal section; a walking beam bearing assembly coupled to the second end of the walking beam; first and second support posts, each extending vertically from the second horizontal section of the support base and coupled to the walking beam bearing assembly, such that the walking beam can pivot with respect to the first and second support posts; first and second support braces, each such support brace extending from one of the first or second support posts to the support base; a pneumatically balanced linear actuator having a lower end and an upper end, the linear actuator being adapted to expand and retract in length and positioned, in a first direction, between the first and second horizontal sections of the support base and, in a second direction perpendicular to the first direction, between the first and second support braces; wherein the upper end of the linear actuator is coupled to the walking beam via an upper actuator bearing, such that the actuator can pivot with respect to the walking beam assembly as the linear actuator expands and retracts in length; and wherein the lower end of the linear actuator is coupled to the support base via a lower actuator bearing, such that the linear actuator can pivot with respect to the support base as the linear actuator expands and contracts in length.

BRIEF DESCRIPTION OF THE DRAWINGS

The following figures form part of the disclosure of inventions and are included to demonstrate further certain aspects of the inventions. The inventions may be better understood by reference to one or more of these figures in combination with the detailed description of certain embodiments presented herein in which:

FIG. 1 illustrates one exemplary embodiment of an improved artificial lift system including a drive system constructed in accordance with the teachings of the present disclosure.

FIGS. 2A and 2B illustrate exemplary positioning of a compressor assembly and, respectively, front and rear views of the assembly of FIG. 1

FIGS. 3A-1 through 3D-2 illustrate aspects of an exemplary actuator assembly and in a process than may be followed to construct the system depicted in FIG. 1.

FIG. 4 illustrates an exemplary apparatus, constructed in accordance with certain teachings of the present disclosure for monitoring the position of the walking beam and/or the horse head and/or any sucker rod (or sucker rod pump) coupled to the horse head in an artificial lift system.

FIG. 5 depicts an exemplary the linear actuator assembly constructed in accordance with certain teachings of the present disclosure that comprises: a drive screw, driven by an electrical motor; a movable ram; and a planetary roller nut that is both engaged with the drive screw and affixed to the movable ram, such that rotation of the drive screw can produce linear movement of the movable ram.

FIGS. 6A-6C illustrates aspects of an exemplary lubrication system that may be used to provide lubrication for components of a linear actuator assembly constructed in accordance with teachings of this disclosure.

FIG. 7 illustrates one exemplary arrangement in which the lubrication system of FIGS. 6A-6C may be used with an angularly articulating linear actuator assembly.

FIGS. 8A-8B illustrate aspects of an alternative system and approach to circulating lubricant (to the lubrication points on a linear actuator assembly.

FIG. 9 illustrates aspects of an exemplary system to both stabilize the counterbalance pressure level at the appropriate level and avoid undesirable over-shots of an artificial lift system constructed in accordance with certain teachings of the present disclosure.

FIG. 10 illustrates pumping unit linking parameters that may be used in implanting certain teachings of the present disclosure.

FIGS. 11A-11B illustrate the relationship between the screw force of the system in relation to the allowable screw force and the screw force of an exemplary system constructed and operated in accordance with certain teachings of the present disclosure.

FIG. 12 illustrates a core mechanism of an exemplary actuator suitable for use in a system constructed in accordance with certain teachings of the present disclosure that comprises a planetary roller screw and nut assembly to drive a linear ram component.

FIGS. 13A and 13B illustrate aspects of an exemplary halting mechanism in the form of a mechanical design having a linear spring component.

FIGS. 14A and 14B illustrate aspects of an alternative exemplary halting mechanism in the form of a mechanical design having a rotary spring primary component.

FIGS. 15A, 15B and 15C illustrate aspects of an alternative exemplary halting mechanism in the form of a design utilizing a hydraulic damping mechanism.

FIGS. 16A-16D illustrate aspects of yet a further alternative halting mechanism that incorporates aspects of both a linear spring and a viscous damper.

FIGS. 17A-17B illustrate the incorporation of a stand-off collar feature into the housing of a centralizer bearing assembly constructed in accordance with certain teachings of the present disclosure.

FIGS. 18 and 19A-19B illustrate details of an alternative exemplary embodiment that includes a planetary roller nut assembly modified to incorporate annular piston/cylinders at each of its end.

FIGS. 20A-20B illustrate characteristics of a low pour point synthetic hydraulic oil during certain exemplary low (artic) temperature conditions.

FIGS. 21A-21C illustrate the accumulator (nitrogen) pressure that may exist in certain embodiments constructed in accordance with the present disclosure, over certain exemplary operating conditions, and the relationship between exemplary piston well and accumulator nitrogen volumes.

While the inventions disclosed herein are susceptible to various modifications and alternative forms, only a few specific embodiments have been shown by way of example in the drawings and are described in more detail below. The figures and detailed descriptions of these embodiments are not intended to limit the breadth or scope of the inventive concepts or the appended claims in any manner. Rather, the figures and detailed written descriptions are provided to illustrate the inventive concepts to a person of ordinary skill in the art and to enable such person to make and use the inventive concepts illustrated and taught by the specific embodiments.

DETAILED DESCRIPTION

The Figures described above, and the written description of specific structures and functions below, are not presented to limit the scope of the inventions disclosed or the scope of the appended claims. Rather, the Figures and written description are provided to teach a person skilled in this art to make and use the inventions for which patent protection is sought.

A person of skill in this art having benefit of this disclosure will understand that the inventions are disclosed and taught herein by reference to specific embodiments, and that these specific embodiments are susceptible to numerous and various modifications and alternative forms without departing from the inventions we possess. For example, and not limitation, a person of skill in this art having benefit of this disclosure will understand that Figures and/or embodiments that use one or more common structures or elements, such as a structure or an element identified by a common reference number, are linked together for all purposes of supporting and enabling our inventions, and that such individual Figures or embodiments are not disparate disclosures. A person of skill in this art having benefit of this disclosure immediately will recognize and understand the various other embodiments of our inventions having one or more of the structures or elements illustrated and/or described in the various linked embodiments. In other words, not all possible embodiments of our inventions are described or illustrated in this application, and one or more of the claims to our inventions may not be directed to a specific, disclosed example. Nonetheless, a person of skill in this art having benefit of this disclosure will understand that the claims are fully supported by the entirety of this disclosure.

Those persons skilled in this art will appreciate that not all features of a commercial embodiment of the inventions are described or shown for the sake of clarity and understanding. Persons of skill in this art will also appreciate that the development of an actual commercial embodiment incorporating aspects of the present inventions will require numerous implementation-specific decisions to achieve the developer's ultimate goal for the commercial embodiment. Such implementation-specific decisions may include, and likely are not limited to, compliance with system-related, business-related, government-related, and other constraints, which may vary by specific implementation, location and from time to time. While a developer's efforts might be complex and time-consuming in an absolute sense, such efforts would be, nevertheless, a routine undertaking for those of skill in this art having benefit of this disclosure.

Further, the use of a singular term, such as, but not limited to, “a,” is not intended as limiting of the number of items. Also, the use of relational terms, such as, but not limited to, “top,” “bottom,” “left,” “right,” “upper,” “lower,” “down,” “up,” “side,” and the like are used in the written description for clarity in specific reference to the Figures and are not intended to limit the scope of the invention or the scope of what is claimed.

Reference throughout this disclosure to “one embodiment,” “an embodiment,” or similar language means that a particular feature, structure, or characteristic described in connection with the embodiment is included in at least one of the many possible embodiments of the present inventions. The terms “including,” “comprising,” “having,” and variations thereof mean “including but not limited to” unless expressly specified otherwise. An enumerated listing of items does not imply that any or all of the items are mutually exclusive and/or mutually inclusive, unless expressly specified otherwise. The terms “a,” “an,” and “the” also refer to “one or more” unless expressly specified otherwise.

The description of elements in each Figure may refer to elements of proceeding Figures. Like numbers refer to like elements in all figures, including alternate embodiments of like elements.

Turning now to several descriptions, with reference to Figures, of particular embodiments incorporating one or more aspects of the disclosed inventions, FIG. 1 illustrates one exemplary embodiment of an improved artificial lift system 100 constructed in accordance with the teachings of the present disclosure.

The illustrated system includes a sucker rod pump assembly (not illustrated) that is connected to a sucker rod 102. The sucker rod 102 may be positioned within a tubing string in fluid communication with a reservoir. The pump and sucker rod 102 are coupled to a drive system for stroking the sucker rod such that the stroking of the sucker rod 102 upwards and downwards will, during normal operation, result in movement of the pump and the displacement of fluids from within the reservoir.

In the illustrated example, the sucker rod pump may be positioned such that at least a portion of the pump may be stroked downwards and upwards within an annulus. In the example of FIG. 1, the annulus corresponds to an interior space defined by a casing string. In the illustrated example, during such strokes, at least a portion of the pump moves within a body of fluid having a fluid. The fluid may take many forms and can be a fluid formed of a mixture of various hydrocarbons, water and/or any other fluid. The pump may, during a single downwards and upwards stroke be fully or partially located within the fluid during the entirety of the stroke.

In the illustrated example, the sucker rod 102 is coupled, through conventional attachment apparatus, to a horse head 104. The horse head 104 is coupled to a walking beam assembly 106. In the example of FIG. 1, the walking beam assembly takes the from of a generally “T” shaped structure. It will be appreciated that the walking beam assembly 106 can be formed from the coupling of different elements or as an integrated structure.

The walking beam assembly 106 is coupled, via Samson post bearings 108, to a Samson post assembly which, in the illustrated example, comprises two upright posts 110a and 110b coupled to each other by a support network. The coupling between the walking beam assembly and the Samson post assembly is such that the walking beam assembly can pivot (about bearings 108) with respect to the Samson post assembly.

As reflected in FIG. 1, in the illustrated example, two Samson post braces 112a and 112b extend, respectively, between the posts 110a and 110b forming the samson post assembly 110 and a support base 114. As further reflected in the figure, a support network, including cross member 116 extends between the Samson post braces 112a and 112b.

In the illustrated example a pneumatically balanced linear actuator assembly 120 is coupled, at a lower end, to the support base 114 via first and second lower actuator bearing assemblies 122a and 122b. In some embodiments, the lower actuator bearing assemblies 122a and 122b may take the form of graphite impregnated bronze bushings. The actuator assembly 120 is, at its upper end, coupled to a bearing saddle 124, via an upper bearing 126. As further reflected in the figure the bearing saddle 124 is coupled to the walking beam 106.

The pneumatically balanced actuator assembly 120 may take the form of the prime mover or drive unit 30 disclosed in U.S. Pat. No. 9,115,574 the disclosure of which is herein incorporated by reference.

In the example of FIG. 1 the actuator assembly 120 is coupled at its lower end to an electrically driven motor 128, which in some examples can be a brushless permanent magnet motor. As reflected in FIG. 1, the support base 114 defines an open interior (not labeled) and at least a portion of the electrically driven motor 128 extends down into the open interior of the support base 114.

In the example of FIG. 1, a lubricating pump assembly 130 is provided for providing lubricating fluids to the actuator assembly 120. As reflected in the figure, in the illustrated example all, or substantially all of the components forming the lubricating pump assembly 130 are positioned within a square region having corners associate with the attachment points between the Samson post elements 110a and 110b and the Samson post braces 112a and 112b.

As noted above, in some examples, the actuator assembly 120 will be a pneumatically balanced actuator assembly. In such embodiments, a compressor assembly 140 can be provided for providing pressurized gas (e.g., air) to the actuator assembly 120. As reflected in FIG. 1, in the illustrated example, the compressor assembly 140 can be positioned at a location outside the described square region having corners associate with the attachment points between the Samson post elements 110a and 110b and the Samson post braces 112a and 112b and the support base 114. Further, in the illustrated example, the compressor assembly 140 is positioned at a location that is generally longitudinally aligned with the actuator assembly 120, and further positioned such that, in the direction perpendicular to a longitudinal length of the walking beam 106, the compressor assembly 140 is contained within a region defined by coupling of the Samson beam posts 110a and 110b to the base 114 and, in some examples, the region defined by the outer edges of the lower actuator bearings 122a and 122b. This positioning of the compressor assembly 140 is further illustrated in FIGS. 2A and 2B which, respectively, present front and rear views of the assembly 100.

In general operation, the assembly 100 will operate as follows: in response to instructions from a system controller (not illustrated) the electric motor 128 can be actuated to cause the length of the linear actuator 120 to extend and retract. As the linear actuator extends and retracts, it will cause the walking beam 106 to pivot about the Samson post bearings 108, thus causing the horse head 104 to raise and lower. This, in turn, will cause the sucker rod 102 to stroke upwards and downwards, moving the rod pump within the well and, during normal operation, displacing fluid from the reservoir.

As may be appreciated, given the geometries of the illustrated system, as the actuator assembly 120 linearly extends and retracts in length, it will pivot both at an upper point where the upper end of the actuator assembly 120 is coupled to the actuator bearing saddle 124 and at the points where the lower end of the actuator assembly 120 is coupled to the lower actuator bearings 122a and 122b.

FIGS. 3A-1 through 3D-2 illustrate aspects of actuator assembly 120 and steps in a process than can be followed to construct the pumping assembly 100 depicted in FIG. 1.

Referring first to FIG. 3A-1, an initial construction point is illustrated showing a status where the Samson posts 110a and 110b and the Samson post braces 112a and 112b have been coupled to each other and to the base 114 and where the respective support networks have been coupled between posts 110a and 110b and posts 112a and 112b. Of note, the cross member 116 is coupled between the support braces 112a and 112b.

As reflected in greater detail in FIG. 3A-2, in the illustrated example, the cross member 116 includes a lug 302, and an adjustable link assembly 304 coupled to lug 302.

FIGS. 3B-1 and 3B-2 illustrate an intermediate actuator construction assembly that includes an actuator assembly 120 to which the following components have been pre-attached: (i) an actuator bearing saddle 124; (ii) a structure including the lower actuator bearings 122a and 122b; and (iii) the electric motor 128. In the illustrated example, the intermediate actuator construction assembly was assembled thorough a process where the actuator assembly 120 was positioned on braces such that it extended longitudinally with respect to the ground (e.g., in position essentially perpendicular to the position illustrated in FIG. 1) such that the connected elements could be conveniently coupled to the actuator. In some embodiments, the intermediate actuator construction assembly can be pre-assembled in a controlled environment (e.g., factory workspace) geographically separate from the location where the assembly of FIG. 3A is assembled.

As further shown in FIG. 3B-1, the illustrated actuator assembly 120 includes three construction coupling points which, in the example, include upper construction connecting elements 310a and 310b and side construction connecting lug 312. As reflected in the figure, each upper construction connecting elements 310a and 310b are located at the end of the actuator assembly 120 closest to the actuator bearing saddle 124 and positioned such that actuator bearing saddle 124 pivots in such a manner that a line passing through the bearing about which the bearing saddle 124 pivots is generally parallel with a line passing through the construction connecting elements 310a and 310b. As further reflected in the illustrated example, the construction lug 312 is located at a portion of the actuator assembly 120 closer to the motor 128 than to the longitudinal midpoint of the actuator 120.

During assembly construction cables, chains, ropes or other elements 325 may be coupled to the construction connecting elements 310a and 310b and the construction connecting lug 312. Such elements may be coupled to a crane or other apparatus to permit movement of the actuator 120. Because of the described arrangement of the construction connecting elements 310a and 310b and the construction connecting lug 312 the actuator assembly 120 may be readily manipulated to be positioned in a horizontally extending direction, as shown in FIG. 3B-1, a vertically extending direction, as shown in FIGS. 3B-2, and directions in between.

As further shown in FIG. 3B-2, the actuator assembly 120 includes lugs 330 and 332, both located in a region of the assembly 120 generally on the side of the assembly 120 opposite the side containing lug 312.

Referring now to FIG. 3C-1, a step in a construction process is illustrated where the pre-assembly depicted in FIGS. 3B-1 and 3B-2 has been positioned adjacent the support assembly depicted in FIG. 3A. As depicted in FIG. 3C-1, because the lower actuator bearings 122a and 12b and the motor 128 had been pre-assembled to the actuator 120, coupling of the pre-assembly can be readily accomplished by lowering the preassembly to a point where the motor 128 extends into an open space within the base 114 and the lower actuator bearings 122a and 122b can be attached to the base 114.

Once the lower actuator bearings 122a and 122b have been attached to the base 114, the actuator assembly 120 can potentially pivot about the lower actuator bearings 122a and 122b. Looking at the figure, it will be appreciated that the extent to which the actuator assembly 120 can pivot in one direction (to the left in the figure) is limited by the cross member 116, which will serve as a hard stop should the assembly 120 pivot in that direction. To limit the ability of the actuator to pivot in the opposite direction (to the right in FIG. 3C-1), the adjustable link 304, that was coupled to the cross member 116, can be coupled to one of the lugs 330a or 330b on the actuator assembly 120. The attachment of the link 304 to one of the identified lugs will serve as a stop preventing pivoting of the actuator assembly 120 in the direction opposite that of the cross member 116. Thus, when the lower actuator bearings 122a and 122b are coupled to the base and the link 304 is coupled between a lug on the cross member and a lug on the actuator 120, the actuator will be positioned such that it is held in a stable position. At that point, the elements 325 can be removed.

In some embodiments, in addition to the adjustable link 304, an additional safety element, such as an operational safety tether, can be coupled between the cross member 116 and the actuator 120. Such an embodiment is reflected in FIG. 3C-2 where an adjustable link 304 is shown coupled between lug 302 of the cross member 316 and the lug 330 of the actuator. and a flexible operational safety tether 318 is coupled between a lower lug on the cross member 116 and lug 332 on the actuator. It will be appreciated that the described tether is exemplary only and that the adjustable link could be coupled between the lower lug on the cross member 116 and the lug 332 on the actuator, with the operational safety teacher 318 being coupled between lugs 302 and 330.

During construction, once the actuator 120 is assembled as shown in FIG. 3C-2, the compressor assembly 140 and the lubricating assembly 130 can be coupled to the actuator assembly and power can be coupled to the system. At this time, the compressor assembly 140 and the lubrication assembly 130 can be activated such that the actuator assembly 120 is balanced and lubricated.

Once the assembly 100 has been partially constructed as described above, a preassembled assembly including the walking beam 106 and the Samson post bearings 108 can be lowered and coupled onto the existing support structure as shown in FIG. 3D-1 and the actuator saddle 124 can be coupled to the walking beam 106 as shown in FIG. 3D-2. In some embodiments, the saddle 124 can be equipped with leverage points (which may take the form of holes) that are configured to receive a force applying member, such as a pry-bar inserted into one of the holes, to aid in slightly pivoting the actuator 120 to assist in lining up the attachment points in the saddle 124 with the corresponding attachment points of the walking beam 106. In such embodiments and/or in alternate embodiments, the adjustable link 304 can be manipulated to make further adjustments of the relative position between the saddle 124 and the walking beam 106. Once the saddle 124 is coupled to the walking beam 106, the horse head 104 can be coupled to the walking beam 106.

Prior to or after the coupling of the horse head 104 to the walking beam, the coupling between the adjustable link 304 and the actuator 120 can be disengaged, such that the actuator 120 is relatively free to pivot within the range limited in one direction by the cross-member 116 and, in the other direction, by the operational safety tether 318.

In the illustrated example, the region of assembly 100 close to the Samsung post bearings 108 is relatively unencumbered such that a novel apparatus for monitoring the position of the walking beam 106 and/or the horse head 104 and/or any sucker rod (or sucker rod pump) coupled to the horse head 104, can be utilized. One example of such an apparatus is shown in FIG. 4.

Referring to FIG. 4, an apparatus is shown that includes a sensor 402 that includes a biased roller element 403 that is biased outwardly but that can be moved inwardly when an inward force is applied to the roller element. The illustrated assembly also includes a Samsung post bearing 108 coupled to an element 400 that pivots and moves as the walking beam 106 pivots about the post bearing 108.

In the illustrated example, as the actuator 120 extends and retracts, and the walking beam 106 pivots about post bearings 108, the element 400 will move such that, at point in time (e.g., when the walking beam is approximately horizontal) the element 400 will generally depicted as shown in FIG. 4 where the region 412 of element 400 is positioned adjacent to the roller element 403 such that the roller element will be full biased outward. As the actuator 120 retracts from the described point, the horse head 104 will move closer to the ground and the element 400 will rotate such that the region 414 is positioned adjacent to the roller element 403 such that the roller element 403 will be inwardly depressed. This inward depression can be used to either actuate a switch or to provide a variable signal proportional to the position of the roller element 403 to provide an indication of the relative position of the walking beam 106 and/or the horse head 104 (or any connected elements).

Similarity, as the actuator 120 extends, and the horse head 104 is elevated, the element 400 will rotate back through a position where region 400 is adjacent the roller element 403 and continue rotating to the point where region 416 of element 400 is adjacent the roller element 403. At that point, the roller element 402 will be depressed and that depression can be used to infer the position of the walking beam 106 and/or the horse head 104.

In the example of FIG. 4, the region 416 of element 400 includes an adjustable stop 417 coupled to the element 400 via an attachment assembly 418 (e.g., a nut and bolt passing through a slot) such that the physical profile of the region 416 can be varied to tune the point at which the roller element 403 is depressed. Note that such an adjustable stop could be included on both sides of the element 400.

The specific form of element 400 in FIG. 4 is exemplary only. For example, while the example element 400 includes a detent at region 414 and an adjustable stop at region 416, detents could be used at both regions. Alternatively, the outer profile of element 400 could be profiled, or stepped, to more precisely indicate the position of the walking beam 106 and/or the horses head 104.

In one exemplary embodiment, the linear actuator assembly 120 may comprise a drive screw, driven by an electrical motor; a movable ram; and a planetary roller nut that is both engaged with the drive screw and affixed to the movable ram, such that rotation of the drive screw can produce linear movement of the movable ram. One such embodiment is reflected by linear actuator assembly 500 in FIG. 5, which depicts a side cutaway view of an exemplary linear actuator assembly.

Referring to FIG. 5 an exemplary linear actuator assembly 500 is shown that includes a drive screw 502 axially and radially affixed at a bottom end by a thrust bearing assembly 504. The drive screw 502 is further coupled, directly or indirectly, at or near, its bottom end to an electrical motor 506, such that rotation of the motor will result in rotation of the drive screw. In the depicted exemplary embodiment, the rotational degree of freedom about the longitudinal axis of the drive screw is generally unrestrained except for its attachment to the motion and, optionally, to one or more brake mechanisms (which may be positioned below the thrust bearing and scal).

The top end of the drive screw 502 is radially affixed with a centralizer bearing assembly whose housing is designed to slide axially inside the bore of a movable machined ram element 510.

In the illustrated example, a planetary roller nut 512 is also positioned within the interior of the bore within the ram 510. In the example, the roller nut 512 is affixed to a lower end of the movable ram 510 and is engaged with the threads of the drive screw 502, such that rotation of the drive screw 502 will result in axial movement of the roller nut and, thus, axial (i.e., linear) movement of the ram 510. Depending on the rotational direction of the drive screw 502, the movement of the ram 510 can be up or down, as reflected by the directional arrows 514.

While the linear actuator assembly 500 of FIG. 5 is beneficial in many respects, such as its compact and robust form, it provides challenges in that the number of movable parts and components warrants regular lubrication to ensure proper functioning and long life of the assembly, preserve seals, cool components, and control potential corrosion. The challenges of lubricating the various moving components are heightened by the fact that many of the moving components, such as the centralizer bearing 508, the planetary roller nut 512, and a section of the drive screw 502, are located within the bore of the moving ram 510. Ideally, a continuous or periodic flow of lubricant (e.g., oil) could be delivered to a point within the bore of the ram 510, such as the point 516.

FIGS. 6A-6C illustrates aspects of a lubrication system that may be used to provide lubrication for components of a linear actuator assembly, such as linear actuator assembly 500. As reflected the figures, the illustrated lubrication system includes a length of oil tubing 602 fluidly coupled to a source of lubricating fluid, such as oil. In the illustrated example, the oil tubing 602 takes the form of a rigid tubular element, although alternate embodiments using flexible or semirigid elements are envisioned.

The oil tubing 602 is coupled at one end to an injection (or spray) nozzle 604 that via a structure that penetrates the outer body of the linear actuator assembly. The injection nozzle 604 is configured and positioned such that the provision of lubricating fluid to the nozzle 604 (via tubing 602) will result in the discharge of lubricating fluid in the direction of the upper arrow in FIG. 6B, such that the lubricating fluid will splash against an outer surface 606 the movable ram 510. As a result of this discharge of lubricating fluid against an outer surface of the movable ram 20) 510, some of, and in many embodiments, the majority of, the sprayed lubricating fluid will cling to the exterior surface 606 of the ram 510 and, due to gravity, begin migrating down the length of the ram 510 in a direction generally indicated by the downward pointing arrows in FIG. 6B. A catch basin or trough 608 is positioned near the lower end of the ram 510 that is configured to catch some or all of the downwardly migrating lubricating fluid and divert the lubricating fluid through one or more openings in the movable ram 510 (one such opening being identified as element 610) such that lubricant flows from the catch basin 608, through one or more openings 610 and is deposited at or near the upper end of the planetary nut 510.

The catch basin 608 may take many forms such as a general bowl shaped structure with one or more openings therein, and the basin 608 may be shaped or formed so as to promote the flow of fluid from the side closest the nozzle 604 to the opposite side. The openings 610 through the ram 510 may be in one or more locations around the surface of the ram 510. For example, embodiments are envisioned wherein there are two openings, on opposite sides. Alternative embodiments are envisioned wherein there are three, four, or more openings. The openings 610 may be of the same diameter or the diameter of the openings may vary to control the flow of lubricant from the basin 608 to the interior of the movable ram 510.

In the illustrated embodiment of FIGS. 6A-6C, the deposition of lubricant on the top of the planetary nut 510 will result in lubrication of the planetary nut 510. Moreover, because the threads of the planetary nut 510 are engaged with the threads of the drive screw 502, the provision of lubrication to the top of the nut 510 will also result in lubrication of the drive screw 502. Because the planetary nut 510 will move along the length of the drive screw 502 during operation of the linear actuator assembly 500, the described system will lubricate the drive screw's threads along all (or substantially all) of the operating length of the screw 502.

Further, in the illustrated embodiment, the coating of the drive screw's 502 threads will result in the slinging of lubricant (through centrifugal motion) off of the screw's threads against the interior surface of the movable ram 510. The movement of the movable ram 510 (and thus the planetary nut 510) will then carry some of the lubrication on the interior surface of the movable ram 510 along with the nut 510, to positions where such lubrication can lubricate the guide rings on the screws upper centralizer bearing assembly 508.

It will be appreciated that lubricating fluid can be constantly provided to the nozzle 604 during operation of the system such that the described components are constantly lubricated. In alternate embodiments, lubrication can be provided to the nozzle 604 according to a duty cycle or a cycle determined by the operating characteristics of the system. Further, the rate at which lubricant is provided to the nozzle can be constant or can vary. In one embodiment, the timing and rate at which lubricant is provided to the nozzle 604 can be linked to the rotational rate of the drive screw 502 such that, during periods of relatively high speed rotations of the screw 502, additional lubricant is provided to the nozzle 604 such that the high speed rotation can result in the discharge of lubricant from the screw 502 against the interior surfaces of the movable ram 510.

In the embodiment of FIGS. 6A-6C, lubricant passing down through the planetary nut 510 is collected in a sump reservoir 612 within the assembly 500. Such a sump 612 may be coupled to a filtration system for redistribution of the filtered lubricant.

As will be appreciated, in the system of 6A-6C lubrication is provided to a number of different components, including portions of a the movable ram 510, by spraying lubricant onto a relatively fixed location of a moving component (ram 510 in the example) and allowing gravitational forces to transport the sprayed lubricant to a pool area (within the basin 608) where it is thereafter distributed to other moving and relatively fixed components (e.g., planetary nut 510, screw 502, and bearing 508).

As described above, in the illustrated example of FIG. 1, the linear actuator assembly 120 articulates angularly during operation of the illustrated system. As such, use of the lubrication system of FIGS. 6A-6C with a linear actuator assembly 120 of the type illustrated in FIG. 1 requires delivery of lubricant to a component that is in motion. FIG. 7 illustrates one exemplary arrangement in which the lubrication system of FIGS. 6A-6C may be used with an angularly articulating linear actuator assembly 120.

Referring to FIG. 7, aspects of pumping system are illustrated that include a skid 114 to which a lubrication pumping assembly 702 is attached. Also attached to the skid 114 is an angularly articulating linear actuator assembly 120, that articulates relative to the skid about a lower bearing assembly 122b.

In the illustrated example, a short length of flexible tubing 706 is used to couple the pumping assembly 702 to a rigid tubing element 602. The ridged tubing element 602 is, in turn, coupled to a nozzle 604 as described above in connection with FIGS. 6A-6C. In the example of FIG. 7, the rigid tubing element 602 is mounted to the angularly articulating assembly 120 such that it articulates with the assembly. The approach of FIG. 7 thus allows the use of an oil circulation system in the form of pumping assembly 702 that is positioned at a fixed location outside of the linear actual assembly 120 (which may have a pressurized interior) coupled to a movable element (in the form of articulating actuator 120). In the illustrated approach the connection between the assembly 702 takes the form of a relatively short flexible hose 706, coupled to a longer ridged component 602, where the rigid component 602, moves with the movable element (assembly 120).

An alternative system and approach to circulating lubricant (sometimes referred to herein as “oil” although other lubricants might be used) to the lubrication points on a linear actuator assembly is depicted in FIG. 8A. This alternative system is suited for various conditions but can be beneficially used under conditions where the ambient temperatures are expected to be mild or warm (30 deg F. to 140 deg F.). This system involves utilizing the pressure variations inside the actuator's counterbalance pressure vessel to pump oil from a lower sump to the desired lubrication points. An example of such pressure variations is depicted in FIG. 8B. These pressure variations are a consequence of the actuator's operation wherein the ram extends and retracts during its operating cycle. Given that the pressure vessel is initially charged with a fixed amount of air, say for example that it is sufficient to produce 400 psig internal pressure at its initial state (which, by convention is assumed to be at the bottom of the stroke). In this initial state, the ram is fully retracted into the actuator vessel and so the compressible volume inside the system is at its minimum. As the ram extends in operation, the interior compressible volume expands and the air pressure declines according to the polytropic compression relationship of an ideal gas. This expression can be written as:

P i := P bot · ( V bot V i ) kair Equation ⁢ 0

Where:

• • kair=specific heat ratio of air (approximately 1.4).
  • P_bot=air pressure inside vessel at the bottom of the stroke (psig).
  • V_bot=interior volume of the vessel-ram system when at the bottom of the stroke (in³)
  • P_i=air pressure inside vessel at the current time (or position) (psig)*
  • V_i=interior volume of the vessel-ram system at the current time (or position) (in³)*
  • The “i” subscript indicates that the value is evaluated at each time step throughout the cycle.

The air pressure inside the vessel (tank) can therefore be expected to vary as graphically indicated above-right with respect to cycle time:

Referring to FIG. 8A, the basic circulation system is comprised of:

• • (i) strainer 802 or other filtering device intended to sequester any captured wear particles and remove them from the lubrication stream,
  • (ii) two check valves 804 and 808 which are oriented to allow fluid motion toward the upper desired lubrication point,
  • (iii) a pre-charged accumulator 806 that will ingest fluid when pressure at its inlet is high and then discharge fluid when pressure is lower,
  • (iv) flow-measuring device 810 such as a flow meter that will indicate the amount of fluid flow proceeding toward the desired lubrication point at all times, and
  • (v) connecting piping and fittings between all devices as indicated in FIG. 8A.

In the example of FIG. 8A, both the entrance (bottom) and exit (top) of this piping string are exposed to the practically the same internal pressure at all times (except for the small static fluid head present at the oil sump entrance). A circulation system that incorporates this lubricating method proceeds as follows:

When the tank pressure is high, the accumulator will ingest oil into its chamber. However, the oil will only come from the lower segment since the lower check valve 804 is the only one allowing flow toward the accumulator 806. The upper check valve 808 is blocked under these conditions.

When the tank pressure is low, the accumulator will expel fluid from its chamber. However, the fluid discharged from the accumulator 806 will only proceed through the upper piping segment since the upper check valve 808 is the only one that is oriented to allow flow away from the accumulator. The lower check valve 804 is blocked during this pressure state.

This alternative method for circulating lubricating oil can be repeated for each of the desired lubrication points. It may also be possible to utilize a common accumulator to feed all of the down-stream lube points, however, it may then be necessary to choke one or more of the fluid flows in order to balance the streams according to the individual lube point needs. In some embodiments, the described approach may be used to “pump” lubricant to a system, such as the one illustrated in FIGS. 6A-6C. Note that, in such an embodiment, if the pressure of the provided lubricant is insufficient to adequately “spray” lubricant through the nozzle 604, some form of a boosting pump could be used. Such boosting pump could potentially be by the reciprocating movement of the pumping system in which the linear actuator is used.

In connection with the described system, it should also be noted that the magnitude of the pressure variation inside the counterbalance pressure vessel is dependent on several factors such as the ram's stroke length and the initial pressure setting (i.e. P_bot). The amount of fluid flow produced by this system will likely decline at shorter stroke lengths or at lower initial pressure values.

In some embodiments of the system described above, the pumping unit may be pneumatically balanced through adjustments of the pressure within a pressure vessel in which the movable ram 510 is located. Approaches for such counterbalancing, which involve the activation of a compressor to increase pressure within the vessel and/or a venting valve to decrease such pressure are set forth in U.S. Pat. No. 10,422,205, the contents of which are incorporated by reference as if fully set forth herein. As set forth in the disclosure of the referenced '205 patent, the provided counterbalance can be controlled in response to imbalances in up vs down stroke motor torque or current by engaging a compressor or venting valve until such time that the torque or current imbalance is reduced to within a prescribed percentage, e.g., 5%.

The approaches of the '205 patent are limited in at least two respects. First, those approaches operate in response to detected motor current and/or torque and not directly in in response to the actual force supplied to the system by the screw actuator. Secondly, at very slow operating rates (very low stokes per minute, or “SPM”), the approach of the '205 patent can result in situations where a pumping cycle may not complete before the compressor or vent valve has overshot its intended target. In such cases a compressor correction may be immediately followed by a venting correction (or vice versa) with neither mode being able to converge the error correction to an acceptable state of being within the 5% accuracy threshold.

To overcome the limitations of the described counterbalancing approach, the balancing corrections can be made based on the screw force and/or can be discretized in such a manner that over-shots be reduced (or eliminated) as described in more detail below. Following the described novel process, a system can be implemented that will both quickly stabilize the counterbalance pressure level at the appropriate level and avoid undesirable over-shots. This can be done by:

• • using the programmed controller processor to compute and record screw force with sufficient resolution to detect peak and minimum values during the cycle regardless of the operating speed of the unit;
  • utilizing the determined peak and minimum screw force values to develop a target pressure inside the counterbalance pressure vessel to correct the imbalance;
  • using the required change in pressure to compute the time required for either the compressor or venting valve to remain engaged and correct the imbalance;
  • operating either the compressor or venting valve for the prescribed time and then disengaging the operation of the element; and
  • reassessing the balance state of the screw to determine if further control loops are necessary.

Aspects of this novel process and system can be understood through a discussion of an example system initially operating under the conditions reflected in FIG. 9.

FIG. 9 illustrates aspects of a system as described above operating under conditions reflected in FIG. 9 which depicts exemplary operating conditions and wherein:

• • the prl curve 904 reflects the polished rod load (lbf);
  • the CBE curve 906 reflects the existing counterbalance effect (lbf);
  • the upper and lower W_p curves 902 reflect the permissible polished rod loads (lbf); and
  • the PRP horizontal axis reflects the Polished rod position (in).

As may be observed by the disparate difference between the upper portion of the prl curve and the upper W_p curve and the lower portion of the prl curve and the lower W_p curve, the pumping unit to which the conditions of FIG. 9 correspond is operating in a slightly underbalanced condition with respect to the permissible loading envelope.

Pursuant to one of embodiment of the present system the appropriate counterbalance correction needed to bring the illustrated system a balanced operating condition may be initiated by initially determining the actuator screw force provided by the system using Equation 1, below:

F screw ⁢ ′ i = ⁢ A · ( pr ⁢ l i + W w ⁢ l ⁢ c ) + W hh . r h ⁢ h ⁢ x i + W w ⁢ b . r w ⁢ b ⁢ x i - I art g · α w ⁢ b ⁢ 2 i - 
 C · F c ⁢ b ⁢ a ⁢ l i · sin ⁢ ( β i ) + C · W r ⁢ a ⁢ s ⁢ s ⁢ y · cos ⁢ ( ϕ - ρ i ) · cos ⁢ ( δ + λ i ) C · sin ⁢ ( β i ) Equation ⁢ 1

Wherein:

• • the pumping unit linkage parameters utilized in the equation are as reflected in FIG. 10;
  • F_screw=axial thrust force exerted by the actuator screw (lb);
  • A=Horizontal dimension from samson post bearing to well head (in);
  • W_wlc=Weight of wireline and carrier bar (lbf);
  • W_hh=Weight of horsehead (lbf);
  • r_hhx=Horizontal dimension from samson post bearing to horsehead center of gravity (in);
  • W_wb=Weight of walking beam (lbf);
  • r_wbx=Horizontal dimension from samson post bearing to walking beam center of gravity (in);
  • I_art=articulating mass moment of inertia of articulating components (walking beam, horsehead, actuator ram assembly, wireline and carrier bar, etc. (in{circumflex over ( )}2-lb);
  • C_wb2=Angular acceleration of walking beam (rad/s{circumflex over ( )}2);
  • C=true distance from the samson post bearing axis to the upper actuator bearing axis (in);
  • F_cbal=force exerted by the actuator resulting from the internal counterbalance pressure (lbf);
  • β=angle between walking beam and actuator ram axis (rad);
  • W_rassy=Weight of the actuator ram assembly (lbf);
  • Φ=Angle of vector from samson post bearing axis to lower actuator bearing axis with respect to the vertical (rad);
  • ρ=Angle of the actuator primary axis with respect to the angle ¢ above (rad);
  • δ=Angle between walking beam flange and the vector C (rad) from above (rad)·
  • g=acceleration of gravity (in/s{circumflex over ( )}2); and
  • the “i” subscript indicates that this term is a variable to be computed at each time step in the operating sequence according to the kinematics of the linkage.

Using Equation 1 the relationship between the screw force of the system in relation to the allowable screw force may be determined.

Examples of exemplary determined and allowable screw forces reflected in FIG. 11A where curve 1102 reflects the minimum and maximum allowable screw force and curve 1104 reflects the screw force of the system during the described operation conditions. As reflected by the differences between the system screw force and the upper and lower maximum allowable screw forces, the illustrated system is operating in an unbalanced condition.

To remedy the out of balance condition in the described conditions, the counterbalance inside the actuator pressure vessel (or actuator jacket) may be increased. To determine the appropriate amount of the pressure increase, in at least some embodiments of the present disclosure, the following process may be followed.

As an initial matter, the basic screw force balancing process (which may be implemented using a programmed processer and an algorithm implanting the approach set forth below) is based on the idea that the peak screw force occurring during the upstroke should be equal in magnitude to the peak screw force occurring during the downstroke. Given that these 2 peak forces will be opposite in sign, this implies that the average of the peak up and down stroke screw force values should be equal to zero.

So as an initial step, a two-point screw force average balance for the example system of FIGS. 10 and 11A, system may be calculated using Equation 2, below:

F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g := max ⁢ ( F screw ⁢ ′ ) + min ⁢ ( F screw ⁢ ′ ) 2 = 2 ⁢ 8 ⁢ 8 9.16 lb ⁢ f Equation ⁢ 2

Because the example system is out of balance, this average is not zero but rather has a positive value.

It has been determined that the internal pressure inside the pressure vessel can be expected to vary during the operating cycle in accordance with the internal compressible volume inside the tank and associated pressure boundary components and that the peak pressure can be anticipated to occur at the bottom of the stroke with the ram fully retracted and where internal volume is at a minimum. For purposes of the presently described embodiment, this maximum pressure will be utilized as the reference key (control point) in subsequent calculations (although it will be appreciated that other control points could be used without departing from the teachings of the present disclosure).

Utilizing the described target points, the next step of the exemplary process will be the determination of a new maximum pressure target for the counterbalance system. In the exemplary embodiment this new maximum pressure target may be set based on the maximum pressure recorded during the previous cycle at the bottom of the stroke (which in the described example of FIGS. 10 and 11A is assumed to be 375 psi) and then augment that value with an incremental pressure difference in the appropriate direction to reduce the pressure error. In this case, owing to the fact that the actuator ram is positioned at very nearly half of the front working center (A dimension) from the samson post bearing, one can use the ratio and the outer diameter of the ram to develop an incremental error correction term in accordance with Equation 3 below:

Imbalance ⁢ Pressure = F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g π 4 · d o ⁢ r ⁢ a ⁢ m 2 Equation ⁢ 3

This term, defined as will be subsequently referred to as the “imbalance pressure”.

Following this approach, the targeted “new” maximum pressure inside the pressure vessel jacket required to balance the screw force can be determined pursuant to FIG. 4, below:

P bot - ⁢ n ⁢ e ⁢ w := P bot + F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g π 4 · d o ⁢ r ⁢ a ⁢ m 2 = 395.92 psi Equation ⁢ 4

Where:

• • P_{bot_new}=the new targeted maximum pressure at the bottom of the stroke (psi);
  • P_bot=the current maximum pressure at the bottom of the stroke (measured during the previous cycle (psi);
  • F_{screw_avg}=The average screw force computed following the previous operating cycle (lbf); and
  • d_oram=Outer diameter of the actuator ram (in).

Considering Equations 3 and 4, the targeted new pressure can then be set as the current peak air pressure inside the pressure vessel jacket plus the current imbalance pressure. This is the new air tank pressure target at the bottom of the stroke (maximum tank pressure). If the new pressure is greater than the current pressure, the compressor should be activated. If the new pressure is less than the current pressure, the venting valve should be activated. These activations should be automatically cancelled once the target pressure has been reached, bearing in mind that the target pressure occurs at the bottom of the stroke (where minimum volume and maximum pressure occur).

Upon adjusting the tank pressure to the new value (using either the compressor or the venting valve as required), the average screw force can be measured/calculated during the subsequent operating cycle using Equation 2 as follows, with a result provided for the present example:

F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g := max ⁢ ( F screw ⁢ ′ ) + min ⁢ ( F screw ⁢ ′ ) 2 = 1 ⁢ 0 8.79 lb ⁢ f

This shows a major reduction in the average screw force. If this value is still over the desired screw force imbalance (say for example 50 lb), The algorithm can be repeated. For example, looking at the projected new target air pressure for the second pass using Equation 4:

P bot - ⁢ n ⁢ e ⁢ w := P bot + F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g π 4 · d o ⁢ r ⁢ a ⁢ m 2 = 396.68 psi

The compressor (or venting valve) can be activated to target the new pressure once again. This time, upon adjusting the tank pressure, the average screw force would drop as follows:

F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g = max ⁢ ( F screw ⁢ ′ ) + min ⁢ ( F screw ⁢ ′ ) 2 = 7.78 lb ⁢ f

Following this approach, for the described example, the continuous graph of the screw force relative to the allowable screw force will be as shown in FIG. 11B, where it can be seen that the adjusted screw forces 1104′ are such that the magnitudes of the peak up and down stroke screw forces are now within 7.78 lb so the unit is balanced within the 50 lb screw force threshold. This described novel balancing process is such that the system can be brought into acceptable balance limits relatively quickly, and in most instances usually be completed within 2-3 iterations.

While the above process sets forth a novel approach for determining the required change in pressure to correct an imbalance condition it does not necessarily resolved the issue of over-correction or overshooting that may occur at low operating speeds. To avoid such over-shoots an additional process may be followed in which the time necessary to raise or lower the internal tank pressure by the specified amount is determined and such determined time may be used to control the system.

Examples of this additional novel process are provided below with respect to two different cases. One in which a compressor may be actuated to increase the pressure within the actuator jacket to correct the screw force imbalance and a second in which a vent valve may be actuated.

Case 1: The Compressor May be Activated to Correct the Screw Force Imbalance.

If the P_{bot_new} value is higher than the current P_bot value and the average screw force is higher than the desired threshold, then the compressor will need to be activated for a period of time to correct the condition. Using numbers from the first example, P_bot=375 psi and P_{bot_new}=395.92 psi for a difference of +20.92 psi, the compressor is selected by the controller for activation.

Based on the ideal gas equations between 2 states, the change in air mass required to raise the pressure by the needed amount can be determined as follows:

STATE 1:

GAGE PRESSURE AT STATE 1	Pg_1 := 375 psi
(This will be the current pressure
at bottom stroke)
ABSOLUTE PRESSURE AT STATE 1	P1 := Pg_1 + 14.7 psi = 389.7 psi
VOLUME IN VESSEL AT STATE 1	V1 := 31.34 ft3
MASS OF AIR PER MOLE	lbmol := 28.97 lb
TEMPERATURE AT STAGE 1 (DEG. F)	TF_1 = 90° F.
ABSOLUTE TEMPERATURE AT	T1 := TF_1 = 549.67 R
STATE 1
(Note: DEG R = DEG F + 450.67)
UNIVERSAL GAS CONSTANT	R u = 1 ⁢ 5 ⁢ 4 ⁢ 5 . 3 ⁢ 4 ⁢ 8 ⁢ 9 ⁢ 6 ⁢ f · lbf lbmol · R
MASS OF AIR CONTAINED IN VESSEL AT STATE 1	n 1 = P 1 · V 1 R u · T 1 = 59.98 lb

In connection with the above, it should be noted that in the example the Gage Pressure at State 1 will be the current pressure at the bottom stroke and that Degrees in R will equal Degrees in Fahrenheit plus 450.67.

Next, the air mass required to achieve the desired target pressure can be determined as follows:

STATE 2:

GAGE PRESSURE AT STATE 2	Pg 2 := 395.92 psi
(This will be the target pressure
at bottom stroke)
ABSOLUTE PRESSURE AT STATE 2	P2 := Pg_2 + 14.7 psi = 419.62 psi
VOLUME IN VESSEL AT STATE 2	V2 := V1
TEMPERATURE AT STAGE 2 (DEG. F)	T2 := T1
MASS OF AIR CONTAINED IN VESSEL AT STATE 2	n 2 = P 2 · V 2 R u · T 2 = 63.2 lb

(Note that in this example both temperature and internal volume are held constant from state 1-2). Computing the change in air mass required permits the determination of air volume required to be added to the actuator jacket under atmospheric conditions:

DIFFERENCE IN MASS FROM STATE 1	Δn := n2 − n1 = 3.22 lb
TO STATE 2
VOLUME OF REQUIRED AIR AT ATMOSPHERIC PRESSURE	SCF := Δ ⁢ n · R u · T 1 14.7 psi = 44.6 ft 3

Utilizing published compressor flow rate data at various output pressures, and/or empirically determined flow rates for the compressor used by the system (and setting the operating pressure at the average of the state 1 and 2 output pressures), a nominal flow rate can be computed through linear interpolation:

COMPRESSOR FLOW RATES AT VAROUS OUTPUT PRESSURES	P samp := ( 200 250 3 ⁢ 0 ⁢ 0 3 ⁢ 5 ⁢ 0 4 ⁢ 0 ⁢ 0 4 ⁢ 5 ⁢ 0 500 ) ⁢ psi ⁢ SCFM samp := ( 18.16 14.52 13.5 12.12 10.8 9.9 8 ) ⁢ ft 3 min
AVERAGE OUTPUT PRESSURE	P op := P 1 + P 2 2 = 4 99.16 psi
BETWEEN
STATE 1 AND
STATE 2
DISCHARGE FLOW RATE OF COMPRESSOR	SCFM op = linterp ⁢ ❘ "\[LeftBracketingBar]" P samp · SCFM samp · P op ❘ "\[RightBracketingBar]" = 19.8 ft 3 min
AT AVERAGE
OUTPUT
PRESSURE

Based on the above, the length of time that the compressor should be activated to achieve the desired internal jacket pressure can be computed:

REQUIRED COMPRESSOR ACTIVATION TIME TO ACHIEVE STATE 2 PRESSURE	t acti ⁢ vation := SCF SCFM op = 2 ⁢ 4 ⁢ 7 . 8 ⁢ 49 ⁢ s ⁢ t activation = 4.131 min	tactivation = 4.131 min

Once the compressor activation time has been fulfilled, the compressor is deactivated.

In some examples, the averaging of the 2 output pressures as described above may. underestimate the necessary compressor activation time. This is believed to be acceptable for control purposes since such a result will lead to a pressure controller undershot rather than an overshot so control loop stability should be preserved.

Case 2: The Venting Valve May be Activated to Correct the Screw Force Imbalance.

If the P_{bot_new} value is lower than the current P_bot value and the average screw force is higher than the desired threshold, then the venting valve will need to be activated for a period of time to correct the condition. Using alternate numbers for illustration, if P_bot=375 psi and the desired target pressure were P_{bot_new}=350 psi for a difference of −25 psi, the venting valve is selected by the controller for activation.

Similar to the above case, the change in air mass required can be determined as follows:

GAGE PRESSURE AT STATE 1	Pg_1 := 375 psi
(This will be the current pressure at bottom stroke)
MASS OF AIR CONTAINED IN VESSEL AT STATE 1	n 1 := P 1 · V 1 R u · T 1 = 59.98 lb
GAGE PRESSURE AT STATE 2	Pg_2 := 3505 psi
(This will be the target pressure at bottom stroke)
MASS OF AIR CONTAINED IN VESSEL AT STATE 2	n 2 := P 2 · V 2 R u · T 2 = 56.13 lb
DIFFERENCE IN MASS FROM STATE 1 TO 2	Δn := n2 − n1 = 3.85 lb

This leads to the ideal gas computation for volume needing to be expelled:

VOLUME OF REQUIRED AIR AT ATMOSPHERIC PRESSURE (SCF)	SCF := Δ ⁢ n · R u · T 1 14.7 psi = - 5 3.3 ft 3

As before, the operating pressure can be taken to be the average of the state 1 and 2 internal pressure and the pressure at the outlet is taken to be at atmospheric conditions:

INLET PRESSURE	Pop = 377.2 psi
OUTLET PRESSURE	Poutlet := 14.7 psi
CHANGE IN PRESSURE BETWEEN INLET AND OUTLET	Δ ⁢ P := min [ ❘ "\[LeftBracketingBar]" P op - P outlet ❘ "\[RightBracketingBar]" - P op 2 ] = 188.6 psi

(NOTE: For many discharge valves, the maximum change in pressure from inlet to outlet cannot exceed a certain percentage of the inlet pressure, e.g. 50% for certain commercially available valves).

Once the desired change in pressure is determined, the governing equation for flow rate through the release valve utilized in the system can be determined using a flow equation provided by the valve supplier (typically in a product catalog) and/or empirical information. For example, for one exemplary valve, the Solenoid Valve Flow Rate (SCFH_) flow rate can be determined as follows:

SCFH - := - [ C V · 960 · Δ ⁢ P psi · ( P o ⁢ p + P outlet ) psi SG · T 1 R ] ⁢ ft 3 hr = - 2337.75 ⁢ ft 3 hr

Where:

VALVE FLOW COEFFICIENT (FROM CATALOG)	CV: = 0.21
SPECIFIC GRAVITY AT S.T.P.	SG: = 1

Finally, the valve activation time duration can be determined as follows:

TIME TAKES FOR VALVE	t act ⁢ _ ⁢ vv := SCF SCFH - = 82.08 S ⁢ t act ⁢ _ ⁢ vv = 1.368 min	tact_vv = 1.368 min
TO
DECREASE
PRESSURE
FROM 375 PSI
TO 350 PSI

Subsequent pumping cycles repeat this control algorithm and affect additional pressure corrections as required.

In the above examples screw force was used to determine the appropriate pressure adjustments required to bring a pumping system as described herein into balance. Alternate embodiments are envisioned wherein such adjustments are based on motor current rather than screw force. This approach has potential benefits because motor current may be sampled directly from sensors. Nevertheless, screw force and motor current are related in a relatively straightforward manner via the following derivation:

Motor current can be written as:

Current M ⁢ ′ i = Tmotor_net i ′ K t

Where:

• • Current_M′=motor current in amps (A)*;
  • T_{motor_net}=motor net torque (ft-lb_f)*; and
  • K_t=rated torque sensitivity (conversion factor from torque to current) ((ft-lb_f)/A).

Motor torque can be written as:

T motor - ⁢ net i = T s ⁢ c ⁢ r ⁢ e ⁢ w i + ( I rotor + I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w ) · α s ⁢ c ⁢ r ⁢ e ⁢ w i

Where:

• • T_screw=screw torque (ft-lb_f)*;
  • I_rotor=mass moment of inertia of the motor rotor component (ft²-lb);
  • I_brake=mass moment of inertia of the brake rotor component (ft²-lb);
  • I_screw=mass moment of inertia of the screw component (ft²-lb); and
  • α_screw=angular acceleration of the screw (radians/s²).

Screw torque can be written as:

T s ⁢ c ⁢ r ⁢ e ⁢ w i = F s ⁢ c ⁢ r ⁢ e ⁢ w i · P h η p

Where:

• • F_screw=Linear screw force (thrust) (lb_f)*;
  • P_h=Screw thread lead (in/revolution)—should be divided by 2*π to yield (in/radian); and
  • η_p=screw practical efficiency.

(NOTE: The “i” subscript in the above equations indicates that the value is evaluated at each time step.)

The above system of equations can be rearranged algebraically to yield the following expression:

F s ⁢ c ⁢ r ⁢ e ⁢ w = η p · [ Current M ⁢ ′ · K t - ( I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w + I rotor ) · α s ⁢ c ⁢ r ⁢ e ⁢ w i ] P h

Where F_screw is now expressed as a function of motor current. Following the previous example above, the average screw force can be computed as:

F s ⁢ c ⁢ r - ⁢ a ⁢ v ⁢ g = max [ η p · [ Current M ⁢ ′ · K t - ( I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w + I rotor ) · α s ⁢ c ⁢ r ⁢ e ⁢ w ] P h ] + min [ η p · [ Current M ⁢ ′ · K t - ( I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w + I rotor ) · α s ⁢ c ⁢ r ⁢ e ⁢ w ] P h ] 2

Consequently, an expression for the desired new counterbalance pressure can be formed:

P bot - ⁢ n ⁢ e ⁢ w = P bot + 
 max [ η p · [ Current M ⁢ ′ · K t - ( I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w + I rotor ) · α s ⁢ c ⁢ r ⁢ e ⁢ w ] P h ] + min [ η p · [ Current M ⁢ ′ · K t - ( I b ⁢ r ⁢ a ⁢ k ⁢ e + I s ⁢ c ⁢ r ⁢ e ⁢ w + I rotor ) · α s ⁢ c ⁢ r ⁢ e ⁢ w ] P h ] π 2 · d o ⁢ r ⁢ a ⁢ m 2

This allows computation of the desired counterbalance pressure using measurements from the drive and controller. From this point the required compressor or venting times can proceed as described above.

One challenge with embodiments using a linear motion actuator, is that-under certain circumstances-such actuators can be places under situations where they may be carrying large inertial forces that could potentially over-run its mechanical stroke limits resulting in impact, lock-up, or damage to the mechanism. These over-run conditions could be the result of:

• • 1. Control logic error-Control system programming or logic is faulty such that an impending crash condition is not detected or recognized.
  • 2. Malfunctioning limit switch-Limit switch positioned at the ends of the linear stroke does not perform as intended.
  • 3. Malfunctioning brake-Brake intended to work in conjunction with limit switch or controller logic malfunctions or is unable to respond in time to prevent the mechanical crash condition.
  • 4. Loss of motor function-Motor intended to drive the actuator suffers a mechanical or electrical failure at or near the ends of the stroke.
  • 5. Operator error-Operator, while operating the actuator in service (manual) mode wherein normal safety protocols are disabled, jogs the actuator beyond its mechanical stroke limit.
  • 6. Load imbalance-Average polished rod load and counterbalance effect forces become grossly unequal due to control system error in the balancing algorithm or other malfunction resulting in a large net force, either upward or downward, acting on the actuator.

In the embodiments of the present disclosure provided herein, such over-run conditions are unlikely to occur due to the multi-layered safety protocols built into the design that are associated with preventing crash conditions. However, stopping the unit in such an event generally depends on the operability of the motor and/or the safety brake components. In the event that one or both of these components are inoperable for any reason, an alternative means of arresting motion is of potential benefit to prevent potentially damaging impact. This alternative arresting component will henceforth be referred to as the “Halting Mechanism” (HM).

List of assumptions at the time of the event:

• • 1. The brake component is off-line and is disengaged (allowing motion).
  • 2. The motor component is off-line and is not controlling motion.
  • 3. At the time of the failure event, the system is operating at the maximum allowed operating speed and is approaching the end of the stroke limit.
  • 4. The pumping unit's mass moments of inertia, both articulating and rotary, are effective and in motion per item 3 above.

Control points: The pumping unit system and associated inertia may be halted in such a way that:

• • 1. A hard crash condition does not occur (i.e. stopping distance) within the actuator assembly.
  • 2. The accelerations (forces and torques) related to halting the pumping unit should generally be controlled such that damage to the working mechanism of the actuator does not occur.

Further Points:

• • 1. The design of the halting mechanism should be compact so that it does not adversely affect the length of the actuator assembly.
  • 2. The halting mechanism, once actuated, should allow the actuator system to be driven back to its nominal home position without further intervention or disassembly once operational safety has been confirmed.
  • 3. Following actuation, the halting mechanism should also return to its ready state without further intervention or disassembly.

The core mechanism of an exemplary actuator suitable for the present discussion is a planetary roller screw and nut assembly that is used to drive a linear ram component as reflected generally in FIG. 12. As reflected in the figure, in the illustrated embodiment, a screw 1202 is constrained axially and radially at its lower end by a thrust bearing 1204 fixedly mounted to a lower bulkhead 1206. The upper end of the screw is constrained radially by a centralizer bearing 1208 deployed in the interior of the linear ram component 1210 but is allowed to slide freely in the linear direction with respect to the ram. The roller nut 1212 is engaged with the threads of the screw 1202 and is fixed to the lower extremity of the ram such that revolution of the screw imparts linear motion on the ram. The ram is constrained to allow only linear motion by a guide tube component 1214 that is in turn fixedly attached to the upper extremity of a pressure vessel 1214. The pressure vessel is fixedly attached to the lower bulkhead 1206 to form an air-tight seal boundary.

Rotary constraint for the actuator body and the ram is provided through connections with the lower actuator bearing assembly 1218 and the upper actuator bearing assembly 1220 respectively. Design of these bearing assemblies provides a degree of freedom allowing the actuator assembly to pivot fore and aft with respect to the pumping unit base and the walking beam. A brake 1222 and motor 1224 are fixedly mounted to the lower bulkhead 1206 and are in turn connected to the shaft of the screw component 1202 to control rotary motion about the screw axis.

The failure scenario described in the sections above could result in a collision involving the roller nut and the thrust bearing at the lower stroke limit at a lower potential collision interface 1226 or the roller nut and centralizer bearing at the upper stoke limit at an upper potential collision interface 1230.

The pumping unit mechanism as a whole is comprised of large and massive articulating components that, when in motion, create large inertial forces. This inertia is a chief element that need be overcome by the halting mechanism (“HM”) to bring the machine to a stop before collision can occur. The design case considered in the following exemplary analysis of the halting mechanism is comprised of the following:

• • 1. Maximum operating speed of the screw: 1100 rpm
  • 2. Screw lead: 50 mm/revolution=1.969 in/revolution
  • 3. Maximum linear speed of the actuator ram: 36.09 in/s
  • 4. System mass moment of inertia referred to the screw shaft axis: 121.86 ft{circumflex over ( )}2*1b
  • 5. Required stopping distance to avoid collision: 3 in
  • 6. Maximum allowable torque on screw-nut interface: 4000 ft-lb

In one exemplary embodiment, the halting mechanism can take the form of a mechanical design having a linear spring component. One example of such an embodiment is reflected in FIG. 13A, with exemplary calculated operating characteristics for such an embodiment being reflected in FIG. 13B. FIG. 13B (and FIGS. 14B, 15B and 16B discussed below) all illustrate the velocity of the ram vs. position to provide a comparative assessment of the halting characteristics of the various halting mechanism embodiments discussed below.

As reflected in FIG. 13A, the illustrated exemplary halting mechanism may include a first spring (or set of springs) 1301 associated with thrust bearing 1304 (coupled to screw 1302) to prevent a collision at one of the potential collision interfaces. The exemplary interface may also include a second spring (or set of springs) 1303 within a ram 1310 to reduce the potential for a collision between a centralizer bearing 1308 and a roller nut 1312.

The linear spring embodiment of FIG. 13A is perhaps the simplest approach for construction of a halting mechanism. It is theoretically able to stop the system in approximately 3.2 inches from initial engagement (within the allowable torque limit) which is slightly further than the desired 3 inches but the required spring rate is approximately 29,810 lbf/in. which yields a maximum spring force of approximately 125.2 kips.

It is noted that the use of a linear spring as the primary halting mechanism component gives rises to various design challenges, at least because the solid height of a linear spring with its heavy coils could consume more than the available design envelope. Other linear spring options such as Bellville washers may be considered.

To address some of the issues posed by the use of a linear spring as the primary halting mechanism component, alternate embodiments may be considered in which a rotary spring is used as the primary halting mechanism component. One example of such an embodiment is reflected in FIG. 14A, with exemplary calculated operating characteristics for such an embodiment being reflected in FIG. 14B. As reflected in FIG. 14A, the disclosed alternate halting mechanism includes a roller screw 1402 coupled to a roller nut 1412 and a thrust bearing 1404. Positioned between the thrust bearing 1404 and the roller nut 412 is a halting structure that includes an elastomeric spring assembly 1456 (that may take the form of one element or multiple elements) positioned between an upper clutch plate 1452 and a lower clutch plate 1454. A cog plate 1450 is also provided with elements that engage the upper clutch plate 1452 such that rotation of the upper clutch plate in one direction will tend to result in compression of the elastomeric spring 1456, inducing a halting force. In the example of FIG. 14B the illustrated halting mechanism is held together by snap ring.

While use of a rotary spring-based halting mechanism as reflected in FIGS. 14A and 14B may be suitable for some applications, for others, the approach may be unsatisfactory primarily due to the design complications associated with accommodating the linear advance of the nut over the 2+ revolutions required to decelerate the system. This embodiment also utilizes linear spring elements deployed along a circumferential path to produce resisting torque. In certain applications, however, this approach may not be able to accommodate a multi-turn deceleration duration which could result in torque overloads in the screw shaft.

In both the linear and coil spring actuated embodiments discussed above the ram will tend to decelerate slowly in the initial stages due to the spring deflection being small. And in such embodiment it is not until later that the spring forces build, and deceleration becomes more abrupt. This leads to rather large torque being applied in downstream components in the latter stages of the deceleration.

Given the character of the speed profiles in the previous spring actuated concepts, yet a further embodiment may be utilized to bleed off kinematic energy early in the deceleration interval. This embodiment involves the use of a type of hydraulic damper mechanism as the main halting mechanism component wherein the maximum resisting torque or force occurs in the region of maximum speed.

One example of such an embodiment, and its calculated operating characteristics, is reflected in FIGS. 15A, 15B and 15C This embodiment utilizes a set of rotating vanes 1560 engaged with a helically grooved sleeve 1562 (with identical pitch to the screw threads to avoid binding) filled with lubricating oil. The vanes 1560 and grooved sleeve 1562 are positioned about the screw element 1502 at a location between the roller nut 1512 and the thrust bearing 1504. When engaged, the vanes 1560 spin down the groove channels in sleeve 1562 attempting to displace the trapped oil. Each vane includes an orifice hole that would allow the fluid to pass through it, albeit with higher velocity. The associated pressure drop caused by the flow restriction exerts a force on the vane that, in turn, yields a resisting torque to slow the rotating screw shaft 1502.

In certain embodiments a restoring spring (or set of springs) 1501 is also provided.

As reflected in FIG. 15C, the deceleration pattern for the embodiment under discussion shows a rapid decrease in speed initially after engagement, when the ram velocity is at its highest, but loses effectiveness as the velocity declines.

Considering the above, yet a further embodiment may be considered that incorporates aspects of both the linear spring and a viscous damper. This embodiment reflects a combination of spring force superimposed with hydraulic damping any, for certain implementations, may provide the most effective solution in halting the actuator motion with the lowest torque or force thereby protecting the internal mechanics of the actuator system.

One example of such an embodiment, and certain calculated operating characteristics, is reflected in FIGS. 16A, 16B, 16C and 16D. As reflected in the referenced figures, this embodiment includes annular piston cylinders (or wells) 1674 and 1675 affixed to the top and bottom faces of the roller nut 1676. In each case, the movable piston (1674 or 1675) faces toward the component with potential to impact the nut 1612, namely the centralizer bearing 1608 on top and the thrust bearing 1604 on the bottom. Each bearing assembly includes a respective stand-off collar (1670 for the centralizer bearing, 1672 for the thrust bearing) fixedly mounted to its housing that will engage with a movable piston (not specifically labeled) in the annular piston cylinder (1675 or 1674) located opposite the collar.

The piston cylinders 1675 and 1674 are filled with suitable hydraulic oil and are connected in tandem with a single nitrogen charged accumulator 1676 creating a hermetically scaled, closed hydraulic circuit.

Upon engagement with its respective stand-off collar, the movable piston will depress, forcing working fluid through one leg of the piping and into the accumulator. The accumulator, being charged with pressurized nitrogen, will exert a resisting pressure against the entering fluid and thus provide a resisting spring force via the displaced piston interface with the stand-off collar.

Additionally, as the fluid passes through the connecting tubing or passages between piston chamber and accumulator, a pressure drop will occur (depending on the speed of the moving piston and the length and diameter of the connecting passages). This creates another component of resisting pressure that acts on the piston and is opposite the direction of motion. This concept therefore provides both a viscous damping force as well as a restoring spring force to slow and ultimately halt the motion of the actuator ram.

The graph in FIG. 16B illustrates the marked difference in the deceleration pattern using this combined concept. Upon engagement, the ram velocity immediately shows deceleration, owing to the viscous pressure losses associated with the hydraulic fluid velocity. Later (as reflected generally in FIG. 16C), the spring effects associated with the building accumulator pressure assert themselves more substantially bringing the ram to a stop at 2.941 inches of total displacement. This is within the 3 inch target displacement.

The piston force exerted to halt the moving ram vs time is shown generally in FIG. 16D. Again, the viscous pressure effect can be seen early in the engagement period followed later by the accumulator spring effect. The maximum piston force of 113.3 kips occurs as the system is finally drawing to a stop and given the low velocity at the time of occurrence, this is almost certainly dominated by the accumulator spring effect. Upon examination of the component pressures acting within the accumulator, it can be seen that the gaseous nitrogen pressure (PN) builds from its initial precharge level of 875 psi to the maximum level of 2847 psi as the system is stopped. On the opposite side of the accumulator boundary, the liquid pressure (P_oilpis) begins initially at 2201 psi due to the high initial piston velocity-viscous effects—but gradually settles back to match the nitrogen pressure as the system velocity goes to zero.

The incorporation of the stand-off collar feature into the housing of the centralizer bearing assembly is reflected generally in FIG. 17A. FIG. 17A reflects an exemplary arrangement in which standoff collar 1670 can be coupled to an exemplary centralizer bearing assembly 1608. Given the arrangement of the centralizer bearing assembly 1608, the piston load path in the example proceeds through the rolling element bearings in route to the screw shaft. This imposes a requirement that the rolling elements inside the bearing assembly 1608 be able to withstand a one-time (or infrequent) load scenario wherein the thrust load exerted on the outer bearing housing, via the stand-off collar, reaches a maximum without damage to the bearing rollers or raceways. Additionally, the means of fastening the centralizer bearing to the roller screw shaft should generally also be capable of withstanding this thrust force without damage. One such exemplary structure is retaining nut 1701 illustrated in FIG. 17A.

FIG. 17B reflects an exemplary embodiment of a lower thrust bearing housing 1703 modified to incorporate a stand-off collar feature 1772 facing the direction of potential impact with the lower piston of the roller nut assembly.

FIGS. 18 and 19A and 19B illustrate further details of this exemplary embodiment. As reflected in FIG. 18, in this embodiment the illustrated planetary roller nut assembly 1812 has been modified in the following ways to incorporate the annular piston/cylinders 1874 and 1875 at each end:

• • 1. The outer housing of the roller nut assembly 1812 has been extended to form the outer walls of the cylinder(s).
  • 2. Threaded sleeves 1881, 1882 have been added to form the interior wall of the cylinders at the end of the assembly. Each threaded sleeve incorporates a static radial o-ring seal to prevent leakage at the threaded interface with the roller nut housing.
  • 3. A single gun-drilled vertical passage 1883 is provided to connect the two piston wells and is intersected at its mid-span with a cross-drilled portm 1884 for connecting the nitrogen charged accumulator 1876 via exterior piping elements. This provides benefit in the reducing the length of connecting passages and potential leak points.
  • 4. Two annular pistons 1874 and 1875 are deployed inside the cylinder wells and are retained via snap rings or similar components to prevent their exiting the cylinder as internal pressure builds upon engagement with the stand-off collar at the opposite end. The pistons are equipped with dynamic elastic seals set at their inner and outer diameters and are designed to slide against the walls of the annular cylinder upon engagement with one or the other stand—of collars.
  • 5. Two ports (1885, 1886) are included to allow initial charging of the hydraulic circuit with oil. A high-pressure check valve (e.g., a Schrader valve) is associated with port 1885 to allow filling the system via hand pump or equivalent device. Port 1886 is provided as a bleeding port to allow air venting during filling but can subsequently be plugged to allow pressurizing the liquid side of the system to a level slightly above the maximum counterbalance air pressure expected inside the pressure vessel. This will ensure that the pistons remain firmly seated against their retainers once placed inside the pressurized environment.

The actuation of the lower annular piston via contact with the stand-off collar is depicted generally in FIGS. 19A and 19B. Upon contact with the annular piston, the stand-off collar will first overcome the static retaining force that otherwise holds the piston firmly against its retaining ring. The magnitude of the retaining force depends on the initial fluid bias pressure inside the piston well which in turn was selected to exceed the maximum expected air pressure inside the pumping unit's pressure vessel by a small amount. The pre-charge nitrogen pressure inside the down-stream accumulator (850 psi in one exemplary embodiment) can be sequestered behind its own partition which is firmly seated at the extent of maximum volume. Once the stand-off collar has exceeded the static retaining force, it need also overcome the pre-charge nitrogen pressure inside the accumulator before annular piston (linear) motion will commence. This therefore constitutes an initial pressure boundary condition inside the piston well just prior to piston movement (P_oilpisinit).

The downward motion of the ram and connected pumping unit mechanism (with associated inertia) depresses the piston into its annular well thereby forcing working fluid through the connecting passages and into the accumulator. The fluid velocity related pressure losses along with the compressing nitrogen work to slow and ultimately stop the motion of the ram as previously discussed.

The actuation of the upper annular piston upon contact with the stand-off collar incorporated into the centralizer bearing is similar in process albeit opposite in direction.

The exemplary pumping units described herein may be designed to work in a wide range of ambient temperature conditions ranging from −40 deg. F. to +140 deg F. One potential complication related to the design of such a unit is changing viscosity of the hydraulic working fluid with temperature. In some environments, typical synthetic lubricating oils may thicken to extraordinarily high viscosities at low temperatures. Such high viscosities can be prone to generating high pressure losses in the passages and piping connecting the cylinder wells with the accumulator. Conversely, at high ambient temperatures, the pressure losses could decline dramatically negating the viscous damping effect that was desirable in the early stages of engagement. To address these issues, in certain embodiments alternate hydraulic oils may be benefically used, including oils with a high viscosity index and a low pour point.

For example, for arctic conditions a low pour point synthetic hydraulic oil having the characteristics reflected in FIG. 20A may be utilized. The performance of such an oil throughout the temperature range can, for certain embodiments, allow the sizing of oil passages and piping to provide the most consistent hydraulic damping response. When such an oil is used, the component pressures of the working fluids inside the accumulator (PN) and at the upstream piston (P_oilpis) at three temperatures spanning the operational range have been calculated to generally correspond to those illustrated in FIG. 20B. As reflected in that figure, the liquid side (piston pressure) may show some effects related to the operating temperature with the initial pressure being higher at lower temperatures. To balance the working fluid pressures between the initial engagement stage (viscous damping) and the latter arresting stage (spring effect) across the applicable temperature range the size of the connecting fluid passages can be appropriately adjusted. If additional refinement of the damping pressure is needed, an orifice can be inserted into the fluid passageways between the piston well and accumulator.

As can be noted from the preceding disclosure, and as reflected in FIGS. 21A and 21B in certain embodiments, the accumulator (nitrogen) pressure begins at the initial pre-charge pressure of 850 psi and climbs in relation to the volumetric displacement created by the annular piston stroke. This proceeds generally according to the polytropic relation for an ideal gas or, more specifically in this instance:

P N = P oilpisinit · V Ninit k N ( V Ninit + A p ⁢ i ⁢ s · x p ⁢ i ⁢ s ) k N

Where:

• • P_oilpisint=initial pressure acting on annular piston.
  • V_Ninit=initial nitrogen volume in accumulator
  • A_pis=area of annular piston
  • x_pis=position of annular piston
  • k_N=specific heat ratio of nitrogen (approximately 1.4)

The initial pressure acting on the annular piston is taken to be equivalent to the combined pre-charge nitrogen pressure in the accumulator plus the residual static retaining force pressure as no motion has yet occurred. The initial nitrogen volume in the accumulator is taken to be total volume capacity of the accumulator. The area and position of the annular piston are calculated values dependent on the geometry and dynamic response of the system.

The volumetric capacity of the accumulator and its initial pre-charge pressure are important parameters to select properly in a system of this type. The volumetric capacity should be large enough to absorb the full swept displacement of the annular piston while leaving sufficient head room inside the accumulator's nitrogen chamber to avoid an exponential pressure spike near the end of the piston stroke. The initial nitrogen pre-charge pressure should be large enough to produce a meaningful spring force even at the beginning engagement stage of displacement.

Other and further embodiments utilizing one or more aspects of the inventions described above can be devised without departing from the spirit of Applicant's invention. Further, the various methods and embodiments of the methods of manufacture and assembly of the system, as well as location specifications, can be included in combination with each other to produce variations of the disclosed methods and embodiments. Discussion of singular elements can include plural elements and vice-versa.

The order of steps can occur in a variety of sequences unless otherwise specifically limited. The various steps described herein can be combined with other steps, interlineated with the stated steps, and/or split into multiple steps. Similarly, elements have been described functionally and can be embodied as separate components or can be combined into components having multiple functions.

The inventions have been described in the context of preferred and other embodiments and not every embodiment of the invention has been described. Obvious modifications and alterations to the described embodiments are available to those of ordinary skill in the art. The disclosed and undisclosed embodiments are not intended to limit or restrict the scope or applicability of the invention conceived of by the Applicants, but rather, in conformity with the patent laws. Applicants intend to protect fully all such modifications and improvements that come within the scope or range of equivalent of the submitted claims.