SCREW PUMP ROTORS AND RING SEALS FOR SCREW PUMP ROTORS

Abstract

A pump rotor for a screw pump, comprising a shaft, a thread on the shaft, the thread comprising a groove disposed on an outer surface thereof, and a seal disposed in the groove. The seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

Description

BACKGROUND

The subject matter disclosed herein generally relates to screw pumps, and, more particularly, to ring seals for screw pump rotors.

In the exploration for oil and gas, the need to transport fluids (oil, water, gas, and foreign solids) from a wellhead to distant processing or storage facilities (instead of building new facilities near the wellheads) is well understood. Twin-screw pumps are increasingly used to aid in the production of these wellhead fluids. The use of the twin-screw pump enables increased production levels by lowering the pressure at the exit of the wellhead as well as a greater total recovery from the reservoir by allowing lower final reservoir pressures before abandoning production.

FIG. 1 illustrates a conventional twin-screw pump 10. This figure is presented simply to illustrate the main components of a twin-screw pump and should not be considered as limiting the invention disclosed herein in any way. As illustrated, the twin-screw pump 10 typically has two rotors 12 and 14 that are disposed within a rotor liner 19. Each rotor has a shaft 18 with one or more outwardly extending sets of screw threads 20 disposed on at least a portion of the length of the shaft 18. The shafts 18 run axially within the rotor liner 19. The two rotors 12, 14 do not touch each other, but their respective opposed screw threads 20 are intertwined. Pump 10 will often be driven by a motor (not shown), which motor rotates rotors 12 and 14. Typically, a drive gear 22 on one of the shafts engages a second gear on the other shaft, such that, when the pump motor turns rotor 12, rotor 14 is turned at the same rate, but in an opposite direction. In operation, wellhead fluids, including particulate materials, are drawn into pump 10 at inlet 24. As the rotors 12 and 14 are turned, rotor chambers 26 formed between adjacent threads 20 displace the wellhead fluids along the rotor shafts 18 towards an outlet chamber 28, which outlet chamber 28 is the point of greatest pressure at the center of the rotors, from where the wellhead fluids are finally discharged from an outlet 30 of the pump 10. The rotor chambers 26 are not completely sealed, but under normal operating conditions the normal clearance spaces that exist between the rotors 12, 14 and between each rotor and the rotor liner 19 are filled with transport fluid. The liquid portion of the transport fluid in these clearance spaces serves to limit the leakage of the pumped fluids between adjacent chambers. The quantity of fluid that escapes from the outlet side of the rotor back toward the inlet represents the pump slip flow, which slip flow is known to decrease the pump volumetric efficiency. As illustrated in FIG. 2 and explained above, pump slip flow (illustrated by the arrows in FIG. 2) can occur between each rotor and the rotor liner 19. Other slip paths include slip between screw tip and adjacent rotors and between faces of the threads 20.

Conventional twin-screw multiphase pumps currently face several challenges. First, assuming a fixed pressure rise per stage, as the total pressure rise requirement increases, the rotor length increases, resulting in an increased rotor deflection under the imposed pressure loading. This deflection creates a more eccentric alignment of the rotors 12, 14 within the rotor liner 19 often resulting in excessive slip between the rotors 12, 14 and the rotor liner 19 or contact and rubbing between the rotors 12, 14 or against the rotor liner 19. Additionally, as the pump slip flow increases, sand particulates trapped in the slip flow can lead to increased erosion or abrasion within the pump, particularly at the rotor tips by a phenomenon referred to as jetting. Such erosion or abrasion can lead to deterioration of the clearance profile and a further increase in the pump slip flow.

It would therefore be desirable to develop a pump rotor that minimizes or eliminates pump slip flow, resulting in a high differential pressure boost multiphase pump with a compact rotor length. In addition, improved sealing between the edges of the rotor and the pump casing will also insure a reduction in solid particulate erosion or abrasion within clearances. It will also be desirable to provide a sealing system that is durable, improves the performance of a pump and does not cause any damage to the pump, even after the sealing system wears.

BRIEF DESCRIPTION

In accordance with one embodiment disclosed herein, a pump rotor for a screw pump comprises a shaft, a thread on the shaft, the thread comprising a groove disposed on an outer surface thereof, and a seal disposed in the groove. The seal is spiraled into the thread from starting point of the thread at an end of the shaft. The groove and the seal are dimensioned to have a clearance between them, enabling the seal to move radially with respect to the thread as the pump rotor is deflected. The groove and the seal form an interlocking mechanism so that the seal is retained in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

In accordance with another embodiment disclosed herein, a method of reducing slip flow in a screw pump having a casing with a low-pressure inlet and a high-pressure outlet, a liner disposed inside of the casing, and a rotor disposed inside of the liner having a shaft and a thread disposed on an outer surface of the shaft, comprises forming a groove on outer surface of the thread and disposing a ring seal in the groove such that the ring seal protrudes outwardly from the groove and rests against an inner surface of the liner of the screw pump to reduce the slip flow from the high-pressure outlet to the low-pressure inlet. The ring seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

In accordance with another embodiment disclosed herein, a twin-screw pump comprises a casing having an inlet and an outlet, a liner disposed inside of the casing and two rotors disposed inside of the liner. Each rotor comprises a shaft, a thread disposed on a portion of an outer surface of the shaft, a groove on an outer surface of the thread and a ring seal in the groove and configured to rotate with the shaft and to protrude outwardly from the groove to rest against an inner surface of the liner. The ring seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

DRAWINGS

These and other features, aspects, and advantages of the present invention will become better understood when the following detailed description is read with reference to the accompanying drawings in which like characters represent like parts throughout the drawings.

FIG. 1 illustrates a conventional twin-screw pump.

FIG. 2 illustrates the pump slip flow path between rotor tips and the liner.

FIG. 3 illustrates a perspective view of a screw pump rotor in accordance with aspects disclosed herein.

FIG. 4 illustrates rectangular cross-sectional ring seal in a groove of a rotor tip.

FIG. 5 illustrates a close-up perspective view of a rotor tip in accordance with an embodiment of the invention.

FIG. 6 illustrates a cross-sectional view of a rotor tip in accordance with an embodiment of the invention.

FIG. 7 illustrates a cross-sectional view of a rotor tip in accordance with another embodiment of the invention.

FIG. 8 illustrates a cross-sectional view of a rotor tip in accordance with another embodiment of the invention.

FIG. 9 illustrates a cross-sectional view of a rotor tip in accordance with another embodiment of the invention.

FIG. 10 illustrates a thread with groove having varying depth in accordance with another embodiment of the invention.

DETAILED DESCRIPTION

Embodiments disclosed herein include screw pump rotors and ring seals for screw pump rotors. The pump rotor comprises a shaft, a thread on the shaft and a ring seal. The thread comprises a groove disposed on the outer surface of the thread and the ring seal is disposed in the groove. The groove and the seal are dimensioned to have a clearance between them to enable the seal to move radially with respect to the thread as the pump rotor is deflected. The seal and the groove are configured such that the seal is retained in the groove while allowing radial displacement of the seal with respect to the thread, as discussed in reference to FIGS. 3, 4 and 5. As used herein, the singular forms “a,” “an” and “the” include plural referents unless the context clearly dictates otherwise.

FIG. 3 illustrates a perspective view of an embodiment of a respective screw pump rotor 40. The threads 44 are helical and are over at least a portion of the shaft 42. The groove 48 is provided on the outer surface or tip 46 of the screw thread 44 that faces the pump liner. The ring seal 50 is typically spiraled into the thread 44 from a starting point of the thread at an end of the shaft 42. The ring seal 50 is helical in structure and may have a length to cover any specific amount of circumferential displacement of the helical threads 44 of the rotor 40. In one embodiment, the ring seal 50 covers one complete revolution of the threads 44.

Pins 60 are used to hold the ring seal 50 in place inside and with respect to the grooves 48 when the rotor is rotated. The pins 60 enable the ring seal to rotate with the shaft. In one embodiment, the ring seals 50 are held in place by the pins 60 disposed once per revolution. In other embodiments, the pins 60 are disposed at any multiple or fraction of revolutions, depending on the circumferential length of the ring seals 50.

As discussed above, the ring seal 50 is important to the overall performance of a screw pump. A clearance between the rotor and the liner is required to allow for rotodynamic vibrations, manufacturing misalignment, and rotor thermal expansion as well as for rotor deflection due to pressure. The ring seal 50 projects out from the threads 44 and is configured to contact the pump liner, filling the clearance between the rotor and the liner. The ring seal 50 of the instant invention provides improved sealing between the rotor and the pump liner. Improved sealing insures a reduction in solid particulate erosion or abrasion within clearances.

Occasionally, the ring seals wear down to an extent where they can get dislodged from the groove. For example, ring seals with rectangular cross-section, as shown in FIG. 4, do not have any constraint and are likely to get dislodged from the groove after some wear which may lead to hard rubs or damage to the rotor or liner.

FIG. 5 illustrates a perspective view of a rotor tip, specifically showing a starting point of the thread 44. As explained above, the ring seal 50 is typically spiraled into the thread 44 from the starting point of the thread.

In one embodiment, the groove 48 and the ring seal 50 comprise an inverted T-shape cross-section, as shown in FIGS. 5 and 6. The groove 48 and the ring seal 50 are dimensioned to have a clearance between them. This clearance provides room for the ring seal 50 to move radially with respect to the thread 44 as the pump rotor is deflected. The inverted T-shape cross-sections of the groove 48 and the ring seal 50 forms an interlocking mechanism that limits the extent of radial displacement of the ring seal 50, as explained below.

The inverted T-shape cross-section of the ring seal 50 can be described as having a first portion 82 substantially parallel to axis 80 of the rotor and a second portion 84 substantially perpendicular to the rotor axis 80 and the first portion 82. Similarly, the inverted T-shape cross-section of the groove 48 can be described as having a first portion 86 substantially parallel to the axis of the rotor 80 and a second portion 88 substantially perpendicular to the rotor axis 80 and the first portion 86. The ring seal, when installed and under normal operating conditions, is designed to spring outward to rest against or adjacent to an inner surface 52 of the pump liner 54, as best shown in FIG. 6. Typically, the second portion 84 of the ring seal contacts the pump liner 54. As the second portion 84 of the ring seal wears out, the ring seal moves radially outward to maintain contact with the pump liner 54. The first portion 82 of the ring seal is displaced radially outward as the second portion 84 of the ring seal wears out. The second portion 84 of ring seal can wear down to an extent until the first portion 86 of the groove prevents the first portion 82 of the ring seal from moving radially outward. Therefore, the T-shape cross-sections of the groove 48 and the ring seal 50 facilitate in limiting the extent of radial displacement of the ring seal 50 and therefore prevent the ring seal 50 from dislodging from the groove 48.

Prior to installation into the thread, the ring seal 50 has a free diameter. During installation of the ring seal 50 into the thread, the diameter of the ring seal is altered, which altered diameter is called a fitted diameter. The contact pressure between the ring seal and the liner is affected by the difference between the free and the fitted diameters of the ring seal. If the free diameter of the ring seal 50 is larger than the diameter of the liner 54, the ring seal 50 needs to be compressed during installation and the contact pressure will be maintained at a higher level. If the free diameter of the ring seal 50 is smaller than the diameter of the liner 54, the contact pressure is reduced or can be negligible until a combination of centrifugal forces and pressure arise to deflect the ring seal 50 outward. Contact pressure will generally decrease with the wear of the ring seal 50, extending life of the ring seal 50.

In operation, as best shown in FIG. 6, as the rotor turns, a profile of increasing pressure develops across the pump. The elimination or minimization of pump slip flow occurring between the rotor 40 and the pump liner 54 is accomplished by an outer surface of the ring seal 50 being pushed against the inside surface 52 of the pump liner. The springing action of the ring seal 50 as well as a centrifugal load on the ring seal 50 caused by the rotation of the rotor 40 pushes the ring seal against the inside surface 52 of the pump liner. A side surface 56 of the ring seal 50 is also pushed against an inner surface 58 of the groove 48 by the pressure difference from one side of the ring seal 50 to the other. The clearance between the ring seal 50 and the groove 48 enables slip flow 90 from beneath the ring seal 50, along the helical groove. Such slip flow 90 from beneath the ring seal flushes sand or other sediment that accumulates in the bottom of the groove 48.

The clearance between the ring seal 50 and the groove 40 enables slip flow 90 from beneath the ring seal 50 and is maintained to provide space under the ring seal free of accumulation without allowing excessive slip flow. The slip flow 90 clears accumulation and allows the ring seal to retract, thereby reducing the contact pressure between the ring seal 50 and the liner 54. Also, in one embodiment, the pins 60 disposed at multiple or fractions of revolutions can be adapted to break the slip flow 90 from beneath the ring seals to control or limit the slip flow 90.

As described earlier, the extent of radial displacement of the ring seal 50 is limited. This form of mechanical restraint will only allow the ring seal 50 to wear down to a point and the remaining part of the ring seal cannot escape the groove. The ring seal 50 is therefore always retained in the groove 48.

In another embodiment as shown in FIG. 7, the ring seal 62 and the groove 64 comprise a cross-section that is a mirror image of an L-shape cross-section. In another embodiment as shown in FIG. 8, the ring seal 66 and the groove 68 are of dovetail-shape cross-section. In both the above embodiments, the ring seals are spiraled into the grooves, the extent of radial displacement of the ring seals 62, 66 is limited, a clearance is maintained between the groove and the ring seal, and the ring seals 62, 66 are prevented from getting dislodged from the grooves 64, 68.

Another embodiment of the ring seal 70 is shown in FIG. 9. In this embodiment, a low-pressure side 72 of the ring seal 70 and a corresponding side 74 of the groove 76 that is facing the low-pressure side 72 of the ring seal are inclined toward the high-pressure side 78 of the ring seal 70. This configuration decreases the contact forces between the ring seal 70 and the inner surface 52 of the pump liner 54, thereby reducing the wear rate of the ring seal 70.

The wear rate of the ring seal can also be reduced by sizing the groove to drop the pressure below the ring seal so that the axial pressure driven component that forces the ring seals against the liner 52, is minimized. The net outward pressure arises because high-pressure fluid leaking under the ring seal forces the ring seal outward. However, because the groove forms a continuous helix about the rotor, extending the groove to a low-pressure inlet and terminating or stopping the groove before it connects to a high-pressure outlet of the pump can relieve the outward pressure. The groove can be cut with varying depth to account for the integration of leakage about the several ring seals and still provide suitable pressure relief, as shown in FIG. 10, where depth of the groove 48 decreases from a depth “d₁” to a depth “d₂”. In one embodiment, the depth “d₁” to depth “d₂” decreases at a constant rate.

When the ring seals 50, 62, 66 and 70 are new, they can substantially seal the gap between rotor 40 and liner 54, even under full deflection and continue to do so after some wear. At some point, the ring seals wear down to an extent where they cannot completely seal the gaps. At this point, the pump performance begins to wane. The performance can be monitored and the worn out ring seals can be replaced with new ring seals. If the ring seals are not replaced at this stage, they may eventually wear down to a point where they are flush with the rotor 40 depending on the degree of eccentricity experienced between the rotor and the liner. If the rotor and liner are always concentric, the excessive wear will be minimal. Although the ring seals provide minimal sealing benefit at this point, they stay in the groove and will not cause any problems with the system.

With respect to the above description, it should be realized that the optimum dimensional relationships for the parts of the invention, to include variations in size, form function and manner of operation, assembly and use, are deemed readily apparent and obvious to those skilled in the art, and therefore, all relationships equivalent to those illustrated in the drawings and described in the specification are intended to be encompassed only by the scope of appended claims.

While only certain features of the invention have been illustrated and described herein, many modifications and changes will occur to those skilled in the art. It is to be understood that the appended claims are intended to cover all such modifications and changes as fall within the true spirit of the invention.

Claims

1. A pump rotor for a screw pump, comprising: a shaft;a helical thread on the shaft, the thread comprising a groove disposed on an outer surface thereof, anda seal disposed in the groove, wherein the seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

2. The pump rotor of claim 1, wherein the seal is a ring seal.

3. The pump rotor of claim 2, wherein the ring seal is configured to protrude outwardly from the groove and to rest against an inner surface of a liner of the screw pump.

4. The pump rotor of claim 3, wherein the ring seal and the groove are configured to enable slip flows from beneath the ring seal.

5. The pump rotor of claim 3, wherein the ring seal is configured to rotate with the shaft.

6. The pump rotor of claim 5, further comprising: a first and a second pins disposed in the groove, wherein the ring seal is disposed between the first and second pins.

7. The pump rotor of claim 6, wherein the ring seal and the thread are helical.

8. The pump rotor of claim 1, further comprising a plurality of pins disposed in the groove and the seal comprises a plurality of ring seals, wherein each of the plurality of ring seals is disposed between a pair of the consecutive pins.

9. The pump rotor of claim 3, wherein the ring seal is a sacrificial wear component of the pump rotor and the ring seal and the groove are configured to prevent the ring seal from getting dislodged from the groove even after the ring seal is worn out.

10. The pump rotor of claim 3, wherein groove and the ring seal comprise an inverted T-shape cross-section.

11. The pump rotor of claim 3, wherein ring seal and the groove comprise a cross-section that is a mirror image of L-shape cross-section.

12. The pump rotor of claim 3, wherein the ring seal and the groove are of dovetail-shape cross-section.

13. The pump rotor of claim 3, wherein a low pressure side of the ring seal and a corresponding side of the groove that is facing the low pressure side of the ring seal are configured to decrease the contact forces between the ring seal and an inner surface of a liner of the screw pump.

14. A method of reducing slip flow in a screw pump having a casing having a low-pressure inlet and a high-pressure outlet, a liner disposed inside of the casing, and a rotor disposed inside of the liner having a shaft and a thread disposed on an outer surface of the shaft, the method comprising: forming a groove on outer surface of the thread; anddisposing a ring seal in the groove such that the ring seal protrudes outwardly from the groove and rests against an inner surface of the liner of the screw pump to reduce the slip flow from the high-pressure outlet to the low-pressure inlet, the ring seal and the groove being configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

15. The method of claim 14, wherein the ring seal and the groove are configured to enable slip flows from beneath the ring seal.

16. The method of claim 14, wherein the groove is formed varying in depth.

17. The method of claim 14, further comprising enabling the ring seal to rotate as the shaft is rotated.

18. A twin screw pump, comprising: a casing having an inlet and an outlet;a liner disposed inside of the casing; andat least two rotors disposed inside of the liner, each rotor comprising, a shaft;a thread disposed on a portion of an outer surface of the shaft;a groove on an outer surface of the thread; anda ring seal in the groove and configured to rotate with the shaft and to protrude outwardly from the groove, wherein the ring seal and the groove are configured to retain the seal in the groove while allowing radial displacement of the seal with respect to the thread as the pump rotor is deflected.

19. The pump of claim 18, wherein the ring seal and the groove are configured to enable slip flows from beneath the ring seal.

20. The pump of claim 18, wherein a low pressure side of the ring seal and a corresponding side of the groove that is facing the low pressure side of the ring seal are configured to decrease the contact forces between the ring seal and an inner surface of a liner of the screw pump.

21. A rotor for a screw pump comprising: a shaft;a helical thread disposed on an outer surface of said shaft, said helical thread defining a groove therein;a seal structure lockingly disposed within said groove, wherein said seal and said groove are configured to allow radial displacement of the seal with respect to the thread as the rotor is deflected.