Patent Grant
8496456
This invention relates to a progressive cavity pump with inner and outer rotors intended for relatively high rotational speeds and great lifting heights with small vibrations.
Progressive cavity pumps, also called Mono pumps, PCP pumps, or Moineau pumps, are a type of displacement pumps which are commercially available in a number of designs for different applications. In particular, these pumps are popular for pumping high-viscosity media. Typically, such pumps include a usually metallic helical rotor which is termed, in what follows, the inner rotor, with Z number of parallel threads which are called thread starts in what follows, Z being any positive integer. The rotor typically runs within a cylinder-shaped stator with a core of an elastic material, a cavity extending axially through it being formed with (Z+1) internal thread starts. The pitch ratio between the stator and rotor should then be (Z+1)/Z, the pitch being defined as the length between adjacent thread crests from the same thread start.
When the geometric design of the threads of the rotor and stator is in accordance with mathematical principles written down by the mathematician Rene Joseph Louis Moineau in, for example, U.S. Pat. No. 1,892,217, the rotor and stator together will form a number of fundamentally discrete cavities by there being, in any section perpendicular to the centre axis of the rotor screw, at least one point of full or approximately full contact between the inner rotor and the stator. The central axis of the rotor will be forced by the stator to have an eccentric position relative to the central axis of the stator. For the rotor to rotate about its own axis within the stator, also the eccentric position of the axis of the rotor will have to rotate about the centre axis of the stator at the same time but in the opposite direction and at a constant centre distance. Therefore, in pumps of this kind there is normally arranged an intermediate shaft with 2 universal joints between the rotor of the pump and the motor driving the pump.
The pumping effect is achieved by said rotational movements bringing the fundamentally discrete cavities between the inner surfaces of the stator and the outer surfaces of the rotor to move from the inlet side of the pump towards the outlet side of the pump during the conveyance of liquid, gas, granulates etc. Characteristically enough, internationally these pumps have therefore often been termed “PCPs” which stands for, in the English language, “Progressive Cavity Pumps”. This is established terminology also in the Norwegian oil industry, for example.
The volumetric efficiency of the pump is determined mainly by the extent to which these fundamentally discrete cavities have been formed in such a way that they actually seal against each other by the relevant rotational speed, pumping medium and differential pressure, or whether there is a certain back-flow because the inner walls of the stator yield elastically or because the stator and rotor are fabricated with a certain clearance between them. To increase the volumetric efficiency, progressive cavity pumps with elastic stators are often constructed with under-dimensioning in the cavity, so that there will be an elastic squeeze fit.
Not very well known and hardly used industrially to any wide extent—yet described already in said U.S. Pat. No. 1,892,217 —are designs of progressive cavity pumps in which a part, like the one termed stator above, is brought to rotate about its own axis in the same direction as the internal rotor. In this case the part with (Z+1) internal thread starts may more correctly be termed an outer rotor. By a fixed speed ratio between the outer rotor and the inner rotor, both the inner rotor and the outer rotor may be mounted in fixed rotary bearings, provided the rotary bearings for the inner rotor have the correct shaft distance or eccentricity measured relative to the central axis of the bearings of the outer rotor.
A limitation to the gaining of ground of such early-described solutions has probably been that an outer rotor needs to be equipped with dynamic seals and rotary bearings, which is avoided completely when a stator is used. On the other hand, an intermediate shaft and universal joints may, in principle, be avoided when the stator is replaced with an outer rotor.
In U.S. Pat. No. 5,407,337 is disclosed a Moineau pump (here called a “helical gear fluid machine”), in which an outer rotor is fixedly supported in a pump casing, an external motor has a fixed axis extending through the external wall of the pump casing parallel with the axis of the outer rotor in a fixed eccentric position relative to it, and the shaft of the motor drives, through a flexible coupling, the inner rotor which has, beyond said coupling, no other support than the walls of the helical cavity of the outer rotor, the material is assumed to be an elastomer. In this case the rotation of the outer rotor is driven exclusively by movements and forces at the contact surfaces of the inner cavity against the inner rotor. A drawback of this solution is that if there is considerable clearance at or elastic deflection of the contact surface, the inner rotor or the outer rotor will be moved more or less away from its ideal relative position. Further, by increasing load, the driving contact surface between the inner and outer rotors will be moved constantly nearer to the motor and force the inner rotor more and more out of parallelism relative to the axis of the outer rotor, so that over the length of the outer rotor, the inner rotor will contact the outer rotor on diametrically opposite sides with consequent friction loss, wear on rotors and motor coupling and also possible signs of wedging. Vibrations, erratic running and reduced efficiency may also be expected.
In U.S. Pat. No. 5,017,087 as well as WO99/22141 inventor John Leisman Sneddon has shown designs of Moineau pumps, in which the outer rotor of the pump is enclosed by and fixedly connected to the rotor of an electromotor whose stator windings are fixedly connected to the pump casing. In these designs the outer and inner rotors of the pump are both fixedly supported radially at both ends in the same pump casing, so that the outer and inner rotors of the pump function together as a mechanical gear, driving the inner rotor at the correct speed relative to the outer rotor which, in turn, is driven by said electromotor. In this case as well, signs of wedging between the inner and outer rotors may arise, in particular if solid hard particles seek to wedge between the inner and outer rotors where these have their driving contact surfaces. Besides, a disadvantage of an inner rotor fixedly supported at both ends is that if the pumping medium is of a kind which must be separated from contact with the bearings, independent dynamic seals will be needed at both ends for both the inner rotor and the outer rotor, as these do not have a common rotary axis.
In U.S. Pat. No. 4,482,305 is shown a pump, flow gauge or similar according to the PCP principle with inner and outer rotors. Here is used a wheel gear outside the pump rotors which ensures a stably correct relative rotational speed between the inner and outer rotors, independently of internal contact surfaces between them. This ensures smoother running, in particular by great pressure differences and/or spacious clearances—which may be necessary to achieve a gradual pressure increase when compressible media are pumped. However, it is assumed here as well that there are dynamic seals and radial bearings at both ends of the inner rotor. The dynamic seal for the outer rotor is also complicated by the diameter of the sealing surface having to be large enough to allow an internal passage for both the pumping medium and the bearing shaft on the extension of the active helical part of the inner rotor.
The invention has for its object to remedy or reduce at least one of the drawbacks of the prior art.
The object is achieved through features which are specified in the description below and in the claims that follow.
The present invention seeks to combine the best aspects of the U.S. Pat. Nos. 4,482,305 and 5,017,087 mentioned, by a wheel gear or similar on one side of the inner and outer pump rotors ensuring smooth and exactly correct relative rotational speed for both rotors independently of the contact surfaces between the inner and outer rotors, while at the same time a bearing, bearing shaft and associated dynamic seal are installed only at the end of the inner rotor at which forces are transmitted from said wheel gear or similar. It is assumed that the outer rotor is made with great flexural and torsional rigidity whereas the inner rotor preferably has great torsional rigidity but little flexural rigidity. These will also be natural properties for the inner rotor considering its other functions. It can be demonstrated that the resultant of the hydraulic forces affecting the inner rotor directly in a radial direction has an approximately constant angular position and, moreover, by constant operating conditions, a constant magnitude. Therefore, the inner rotor will tend always to lean towards the same side, where it will get support at all times from the cavity walls of the outer rotor at a number of contact surfaces corresponding to Z× the number of revolutions of the screw of the inner rotor. These contact surfaces will move linearly towards the outlet side of the pump and be renewed for every revolution. This gives very moderate vibrations. The length and flexural rigidity of the supporting shaft of the inner rotor and clearances between the inner and outer rotors can easily be adjusted in such a way that changing bending stresses caused during operation, by the inner rotor leaning against the supporting surfaces of the outer rotor, will be acceptable. The principal stresses on the inner rotor will be approximately constant torsional stresses.
Please note that in contrast to U.S. Pat. No. 5,407,337 all the contact surfaces against the outer rotor will be on the same side in a PCP pump according to the invention, so that at great pressure differences and/or clearances between the rotors, the active helical part of the shaft of the inner rotor will tend towards a minor parallel translation instead of an angular displacement, so that bearing forces will not occur on opposite sides of the inner rotor. An extended axis towards a bearing will have a deflection, whereas it can be demonstrated that twisting in consequence of torsion will usually be moderate. When an undesired hard particle positions itself between the contact surfaces of the inner and outer rotors in a PCP pump according to the invention, the driving torque applied to both the inner rotor and the outer rotor from the wheel gear will tend to push the particle away or let the particle roll between the contact surfaces. This different from the situation when, for example, the inner rotor is driven only by the contact surfaces against the outer rotor—as in U.S. Pat. No. 5,017,087 and WO99/22141—in which such a particle will reduce the torque arm of the driving force and may even turn its direction.
In a PCP construction according to the invention and with the wheel gear placed on the inlet side, there will only be a need for one dynamic seal on the outlet side, at which the pressure is the greatest. It seals against the outer rotor. In a preferred embodiment, the diameter of the contact surfaces of the seal can be minimized by the inner rotor being terminated upstream relative to the seal so that its area will not make a deduction from the effective flow area of the outer rotor, by the necessary effective flow area being gradually changed into a circular shape up to the position of the seal, and by the diameter of the seal exceeding the diameter of the flow area to the least extent possible. A mechanical seal in which both a spring-loaded part and one abutment seat are arranged for tight internal mounting in bores will be a preferred embodiment here.
A progressive cavity pump comprising at least one inner rotor with Z external threads and at least one adapted outer rotor with Z+1 internal threads is thus characterized by the outer rotor having at least two radial bearings—preferably one close to either end—whereas the inner rotor has a radial bearing only to one side of its helical part, and by there being arranged, on the same side as the bearing of the inner rotor, a conventional gear, for example a wheel gear, which is arranged to maintain a stable ratio between the rotational speeds of the inner and outer rotors, equalling the ratio (Z+1)/Z independently of driving contact between the helical surfaces of the inner and outer rotors.
The progressive cavity pump may be formed in such a way that the conventional gear and the bearing of the inner rotor are arranged on the inlet side of the pump.
The progressive cavity pump may be formed in such a way that the diameter of the dynamic seal of the outer rotor on the outlet side is minimized by the inner rotor being terminated upstream relative to the seal, that a flow area in the outer rotor is circular under a seal and that the cavity cross section of the outer rotor is reduced in this area, in principle according to the cross-sectional area of the helical part of the inner rotor.
The progressive cavity pump may be formed in such a way that as a seal on the outlet side of the outer rotor is used a mechanical seal of such design that both static and dynamic parts are adapted for installation internally in bores.
In particular for installation in narrow pipes and where low reservoir pressure on the suction side gives a risk of cavitation, the progressive cavity pump may be formed in such a way that the conventional gear and the bearing of the inner rotor are arranged on the outlet side of the pump, that the inner rotor starts downstream relative to the bearings and dynamic seals of the outer rotor on the inlet side, and that the flow area upstream relative to the helical parts of the inner and outer rotors is circular with a maximal area relative to the space available within the bearings and seals of the outer rotor.
The progressive cavity pump may be formed in such a way that a wheel gear is used, in which an external toothing on the inner rotor is engaged with an internal toothing on an intermediate wheel having its rotary axis parallel with the inner rotor and eccentrically on the opposite side relative to the axis of the outer rotor, the intermediate wheel also having an external toothing, and the external toothing of the intermediate wheel being engaged with an internal toothing on the outer rotor.
The progressive cavity pump may be formed in such a way that the motor is arranged eccentrically relative to the outer rotor for and drives the inner rotor directly via a conventional coupling.
The progressive cavity pump may be formed in such a way that the motor is arranged concentrically and drives the inner rotor via an intermediate shaft with two universal joints.
The progressive cavity pump may be formed in such a way that the outer rotor is driven directly by a motor by the rotor of the motor being fixedly connected to and concentrically surrounding the outer rotor of the pump, and the stator of the motor being fixed in the same housing as the bearings of the outer rotor of the pump, said housing consisting of one part or possibly several parts rigidly connected to each other.
The progressive cavity pump may be formed in such a way that the motor is installed concentrically with the axis of the outer rotor, and that on the drive shaft of the motor is mounted a gearwheel engaged with a gearwheel on the inner rotor of the pump.
The progressive cavity pump may be formed in such a way that the rotor of the motor constitutes a direct extension of the outer pump rotor on the opposite side of the conventional gear, that the rotor of the motor only or partially has the same bearings and dynamic seals as the outer pump rotor, and that the rotor of the motor has an internal, preferably circular cavity in the direct extension of the helical cavity of the outer pump rotor.
The progressive cavity pump may be formed in such a way that the fixed connection between the outer pump rotor and the rotor of the motor contains substantially radial openings which allow parts of the pumping medium to flow externally past the motor to contribute to the cooling thereof.
The progressive cavity pump may be formed in such a way that dynamic seals are arranged on both sides of said openings to prevent the pumping medium from direct contact with motor windings or bearings.
The progressive cavity pump may be formed in such a way that the rotor of the motor is hollow, allowing flow-through of pumping medium.
The progressive cavity pump may be formed in such a way that there is used a wheel gear comprising for the outer rotor a driving gearwheel with external toothing surrounding the cavity of the outer rotor, for the inner rotor a driving gearwheel with external toothing surrounding a flow area, and for the pump motor a driving gearwheel arranged concentrically relative to the outer rotor, that there are at least two planet shafts with rotary axes arranged at the same fixed distance from the rotary axes of the motor and outer rotor, that each of the planet shafts contains a respective gearwheel for constant engagement with the gearwheels of the outer rotor and the motor, respectively, and that one or two of said planet shafts additionally contain(s) a gearwheel for constant engagement with the gearwheel of the inner rotor.
The progressive cavity pump may be formed in such a way that the motor surrounds the outer rotor in such a manner that the outer rotor and the rotor of the motor are combined and rotate together in common bearings, that on one side—preferably the inlet side—the outer rotor has an external toothing which is constantly engaged with one or two planet wheels, each with two wheels on a common shaft, one of which is engaged with the gearwheel of the outer rotor and the other is engaged with a gearwheel fixedly and concentrically mounted on the inner rotor, and that the gearwheels together form the gear ratio (Z+1)/Z between, respectively, the inner rotor and the outer rotor of the pump.
In what follows in described an example of a preferred embodiment which is visualized in the accompanying drawings, in which:
In the drawings, the reference numeral 1 indicates in
The liquid intake of the pump occurs in this case between the pump sections 6 and 7, so that on its way to the suction intake the crude oil flows externally past the motor 8, contributing to the cooling thereof.
It should be noted that even though in this exemplary embodiment the pump is assembled from sections screwed together is axially, completely different embodiments will also be possible. For example, an embodiment is conceivable in which the entire pump casing is split into 2 longitudinal parts, each part extending over the full length from the motor to the outlet flange, but in which the dividing line between the parts substantially follows a plane through the central axes of the inner and outer rotors.
From the section in
To reduce the risk of wedging or substantial vibrations between the inner and outer rotors, it is assumed according to the invention that a wheel gear or similar is arranged, see
A lubricating and cooling agent—preferably also used as the fluid in hydrodynamic bearings—is supplied in small dosed amounts, intermittently or continuously, through a small pipe connection in an intake opening 31, filling the cavity surrounding a flexible pressure-equalizing diaphragm 28. The space on the inside of the pressure-balancing diaphragm has an open connection to a static outlet cavity 30, but is otherwise completely tight. Even by rather approximate dosing of the lubricating and cooling agent—for example controlled by signals from a pressure sensor—the pressure difference across the mechanical seal 27 will stay close to zero at all times, so that the leakage will be very small and the mechanical bearing pressure on the sealing surface may be limited in favour of low friction.
In order to flow on upstream within a cavity 25 to a bearing 24 etc., the lubricant filling a cavity 29 outside the seal 27 at a pressure corresponding to the outlet pressure must pass narrow fits in the labyrinth seal 26 which may be assembled from several sections of a conventional type or, for example, be made with helical grooves between annular grooves, so that during rotation the helical grooves will bring about increased counterpressure.
The function of the labyrinth seal is not primarily to retain contaminants but to bring about a substantially lower pressure in the cooling and lubricating agent on the inlet side than on the outlet side. The dynamic seals, for example in the form of conventional wearing rings which will possibly be needed to separate the pumping medium from the lubricating and cooling liquid at both wear rings 34a, 34b and an outer rotor 34c, 34d on the inlet side, may therefore have a moderate external overpressure controlled by the flow rate applied through an intake 31 and the labyrinth seal 26. This flow rate may either be identical to an accepted leakage flow through the wearing rings—which let some lubricating and cooling agent get mixed into the crude oil and be recovered together with it—or be greater, by the excess parts being conveyed in separate return lines, not shown, back to a lubricating and cooling agent feed pump. In the example there are further wearing rings 34e, 34f which prevent leakage of lubricating and cooling liquid along the motor shaft 9. In the cavity 20 between the outer rotor and the external walls of the pump casing there may be advantageously arranged a cooling circulation of cooling and lubricating agent, for example by the outer rotor being made with helical external grooves, wings, or similar, not shown.
From
The transition between the drive shaft 35 and the extension 10 of the inner rotor with the bearing 11 and the gearwheel 33 is formed, for reasons of installation, by a releasable connection, for example in the form of splines 35a and a central bolt which is screwed from a hollow within the toothing 32 into the drive shaft 35. The wear ring 34a is mounted on a flexurally rigid cylinder surface and has a wear coating with a precision fit adapted to a wearing ring 34b assumed to be mounted in section 6 of the pump casing. Together the wearing surface 34a and the wearing ring 34b form sufficient sealing between the space filled with lubricating oil surrounding the gearwheel 33 and the pumping medium, for example crude oil with a certain content of sand particles, surrounding the sections 35, 36, 18 and 37 during operation. On account of precision in the wearing ring and tooth engagement, the hydrostatic bearing 11 shown may possibly be replaced with a particularly torsionally rigid and positionally accurate roller bearing or bearings mounted in pairs.
In a gear embodiment which is shown with parts partially cut through in
In the exemplary embodiment shown this is achieved by the gearwheel 33 having z1=33 teeth, the gearwheel 39 having z2=57 teeth, the gearwheel 40 having z3=76 teeth and the gearwheel 41 having z4=88 teeth. The gear ratio then meets the requirement by:
(z3/z4)*(z1/z2)=(76*33)/(88*57)=½=Z/(Z+1)
To make it possible to mount the motor shaft 9 and gearwheel 38 concentrically relative to the outer rotor 41, 42, 19 etc.—without the use of an intermediate shaft with cardan joints, which is usual in classic PCP pumps—the difference between the tooth number z5 of the gearwheel 32 on the extension of the inner rotor and the tooth number z6 of the gearwheel 38 on the motor shaft must be:
(z5−z6)˜±(2*E)/m1
in which E is the eccentricity or distance between the central axes of the inner rotor and the outer rotor, or motor, m1 is the gearwheel module of this pair of teeth, and a certain degree of approximation is allowed by involute toothing, for example. This requirement does not normally constitute a critical limitation as the motor may have rotational speed control (VFD), as the gear ratio between the motor and pump rotors can be varied by varying the tooth numbers z5 and z6 in parallel, and as the gear ratio may be made greater or smaller than 1 by switching any of the gearwheels 38 and 32 which are given internal toothing and external toothing, respectively.
However, what is more limiting to selectable combinations of module and tooth number is the following requirement:
(z4−z3)˜(2*Em)/m2
in which Em is the shaft distance between the outer rotor—coinciding with the gearwheel 41—and the intermediate wheel 13, and m2 is the module of this pair of teeth, which may be different from the module of the pair of teeth 32, 38. It is assumed here that the centre axis of the intermediate wheel is in the same plane as the axes of the outer and inner rotors, but that the intermediate wheel is on the opposite side relative to the inner rotor.
It is further required that:
(z2−z1)˜(2*(E+Em))/m3
in which m3 may be a further module for this pair of teeth. The selectability of m1, m2, m3 and Em will in most cases make it possible to adapt a suitable gear according to the principles of the exemplary embodiment shown. In the figure examples used, the requirements are met with equal gearwheel modules and opposite equal eccentricities:
m1=m2=m3 and Em=E.
In
(zo/zP1)*(zP2/z1)=(Z+1)/Z
in which Z is the number of thread starts of the inner rotor Z=1 of
In
The conventional gear which is to maintain, according to an embodiment of the present invention, the correct relative speed of rotation between the inner and outer pump rotors independently of driving contact directly between their helical surfaces is formed, in this case, by a substantially simplified version of the gear shown in
(z143/z148)*(z147/z133)=(Z+1)/Z=2
in which z143, z148, z147, z133 are the tooth numbers of 143, 148, 147 and 133, respectively.
| Number | Date | Country | Kind |
|---|---|---|---|
| 20083617 | Aug 2008 | NO | national |
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/NO2009/000276 | 8/7/2009 | WO | 00 | 2/17/2011 |
| Publishing Document | Publishing Date | Country | Kind |
|---|---|---|---|
| WO2010/021550 | 2/25/2010 | WO | A |
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| Number | Date | Country | |
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| 20110150687 A1 | Jun 2011 | US |
