WO2017175088A1

WO2017175088A1 - Twin-screw positive displacement pump

Info

Publication number: WO2017175088A1
Application number: PCT/IB2017/051707
Authority: WO
Inventors: Antonio CHESSA
Original assignee: Chessa Antonio
Priority date: 2016-04-05
Filing date: 2017-03-24
Publication date: 2017-10-12
Also published as: ITUA20162324A1

Abstract

A positive displacement pump (1) comprising at least a first (2) and a second (3) rotor at least partially sealingly accommodated inside a pump body (4), the first and second rotor having synchronization means (5) of their rotation, each rotor comprising a first (2A, 3A) and a second (2B, 3B) screw pumping profile between which a central part (2C, 3C) without pumping profile is interposed, the first (2A) and the second (2B) profile of the first rotor (2) being engaged with the first (3A) and the second (3B) profile of the second rotor (3), respectively, so as to pump a fluid processable by the pump by sucking it from a first (6A) and a second (6B) suction associated to the end sides of the rotors (2, 3) and pumping it towards said central part (2C, 3C) to which a delivery (7) of the pump is associated; the central part (2C, 3C) of each rotor is coupled to a hydrodynamic bearing (8) accommodated inside the pump body (4) operated by the same fluid processed by the pump.

Description

TWIN-SCREW POSITIVE DISPLACEMENT PUMP DESCRIPTION

FIELD OF THE INVENTION

The present invention relates to a twin-screw positive displacement pump.

In particular, it relates to a twin-screw positive displacement pump, specifically for the high pressure pumping of water or other types of fluid.

PRIOR ART

As is known, twin-screw pumps are widely used in several fields of the industry.

Some types of positive displacement pumps are provided with a pair of rotors in synchronized rotation, each comprising a pair of screw profiles separate by a part without profile located at the centre of the rotor and of variable size.

When the two rotors are engaged and set in rotation, the fluid trapped between the screws is sucked by the ends of the rotors and conveyed under pressure towards the central part of the rotors, where a pump delivery is normally present.

Known twin-screw pumps do not work in optimal conditions when they are used to process fluids at high pressures.

DE 102005 037118 B3, describes a twin-screw pump with a hydrodynamic bearing at the centre of the rotors. The oil used to operate the bearing may leak through gaskets and contaminate the fluid processed by the pump.

SUMMARY OF THE INVENTION The object of the present invention is to provide a twin-screw positive displacement pump which is improved compared to conventional pumps.

This and other objects are achieved by a twin-screw positive displacement pump implemented according to the technical teachings of the appended claims. An advantage of the pump according to the invention consists in its ability to carry fluids at high pressures in a reliable and cost-effective manner.

A further advantage of the pump according to the present invention is that it allows pumping low viscosity fluids, such as water.

BRIEF DESCRIPTION OF THE FIGURES Further features and advantages of the invention will become apparent from the description of a preferred but non-exclusive embodiment of the device, shown by way of a non-limiting example in the accompanying drawings, in which: figure 1 shows a simplified transverse section taken along line I-I in figure 2 of a pump according to the present invention; figure 2 is a simplified transverse section taken along line II-II in figure 1 ; figure 3 is a simplified section taken along line III-III in figure 1 ; figure 4 is a simplified section taken along line IV-IV in figure 1 ; figure 5 is a simplified section taken along line V-V in figure 1 ; figure 6 shows a simplified section of a detail of the pump in figure 1 ; figure 7 and figure 8 are enlarged, simplified, and out of scale sections of a detail of the pump in figure 1 ; figure 9 shows a simplified and partial section of an engaged part of the rotors of the pump according to the present invention; figure 10 is a simplified sectional view of a tooth of the rotors in figure 9; figures 11, 12 and 13 are enlarged views of the parts enclosed in the circles in figure 9; and figure 14 is a perspective view of two rotors and a partially exploded support bearing of the rotors.

DETAILED DESCRIPTION OF THE INVENTION

With reference to the above figures, a twin-screw positive displacement pump is shown, indicated as a whole with reference numeral 1. The twin-screw positive displacement pump 1 is shown in a sectional view in figure 1. From the section, it is well seen that it comprises at least a first 2 and a second 3 rotor which are at least partially sealingly accommodated inside a pump body 4.

The rotors are advantageously symmetrical and can be mounted in the pump body in a reversible and interchangeable (i.e. positioning) manner, allowing an adaptation of the pump to flexibility of installation of the same both on already operating systems and in the design stage.

The pump body 4 may be formed by a tubular elongate hollow body 4A, internally shaped and closed at the two ends by two heads 4B. Bearing are mounted on the heads, for example of radial type, intended to support the two rotors. The two heads 4B are fixed to the elongated hollow body 4 A through appropriate screws 50. As can be seen in the drawing, the two heads 4B, in addition to the bearings, are provided with sealing means 51 intended to cooperate with the rotors so that they are sealingly accommodated in the pump body 4.

It can be seen that a part of the rotors is accommodated right inside the pump body 4, while a part of them visible to the left of figure 1 can protrude from the latter. Specifically, the protruding part of the shafts comprises a pair of synchronization means of the rotation of rotors 2, 3, which may comprise gears 53. Gears 53 are advantageously of the helical herringbone type (preferably stopped by locking assemblies with automatic compensation of the axial thrust) so as to minimize or cancel the residual axial thrust on the rotors and help maintain the latter in their design position. Such a feature is extremely advantageous for the operation of the pump, since the thrusts generated by the motion synchronization means of the two rotors are compensated and are not relieved on the rotors themselves nor on the bearings that support the rotors.

The high efficiency synchronous gears with helical opposing teeth shape ensure the total (or almost total) absence of any axial thrust, thus placing the rotors in a well-defined position and preventing them from moving from the design position in any operating condition. There is therefore no possibility to have vibrations of the rotors and friction between them, or such a possibility is absolutely small. The particular shape of the arrow gears also ensures a high mechanical efficiency, with reduction of the heat to be dissipated outside and at the same time allows particularly low noise levels.

The self-centring locking assemblies with compensated or substantially compensated axial thrust allow safe and correct assembly and quick and easy disassembly of the gears. The gears are always mounted in the design position for an optimal functioning thereof. This is allowed by the possibility provided by the locking assemblies to adjust the position of a pair with respect to the other.

On the opposite side of gears 53, another part of the second rotor 3 projects from the pump body, and specifically a portion of the driving rotor 3 intended to be torsionally associated in a known manner, for example by a key housed in a suitable seat 54, to a pump driving means (not shown). The driving means of the driving rotor may be for example an asynchronous electric motor.

It can be seen that, casings 56 may be provided, associated to heads 4B of the pump body, intended to protect gears 53 or in any case to protect the end portions of the rotors possibly protruding from the pump body 4. The first 2 and the second 3 rotor are each provided with a first 2A, 3 A and a second 2B, 3B worm screw pumping profile. They may be made in one piece on the rotors by chip removal.

The first and the second pumping profile of each rotor are spaced by a central part 2C, 3C without pumping profile, which is interposed between the profiles and has a substantially circular cross section.

For the pumping action to take place, the first 2 A and the second 2B profile of the first rotor 2 are engaged with the first 3A and the second 3B profile, respectively, of the second rotor 3. In this way, when the rotors are in rotation, a fluid processable by the pump is sucked by a first 6A and a second 6B suction associated to the end sides of rotors 2, 3 neighbouring the pumping profile and is pumped between the two screws of the central part 2C, 3C to which a delivery 7 is associated (fig. 2).

According to an aspect of the present invention, the central part 2C, 3C of each rotor is coupled to a bearing 8, preferably of the hydrodynamic type, accommodated inside the pump body 4.

In fact, bearing 8 acts on the central part of each rotor, supporting it and minimizing the bending thereof, in particular during the pump operation.

Advantageously, bearing 8 is of the hydrodynamic type and comprises at least one inner cylindrical wall 810, hydrodynamically intended to cooperate with said central part 2B, 2C of each rotor to support it, and an outer perimeter wall 811 which can be associated to the pump body 4, for example through a plurality of screws 60 simply clamped against the outer perimeter wall 811 (fig. 6) of the bearing.

It is seen that bearing 8 is provided, at the ends thereof facing the pumping profiles of each rotor 2, 3, with a chamber 10A, 10B (see fig. 4, one for each rotor) intended to accommodate at least part of the pumped pressurised fluid. The chamber is for example shaped as two opposing C. Another part of the pumped fluid is wedged between bearing 8 and the central part 2B, 2C of the rotors. Chamber 10A, 10B is in communication with delivery 7, as seen in figure 2.

Figures 7 and 8 show, in particular, that each inner cylindrical wall 810 associated to the central part of the respective rotor is eccentric (by a distance "e") with respect to the latter, and has a slightly greater diameter than that of said central part 2B, 2C. By virtue of this eccentricity, when the rotor is in rotation it generates a hydrodynamic thrust I

(fig. 8, which is very simplified and purely indicative) which suitably supports the rotors.

As is seen in figures 7 and 8, the centre of the central part CPC of the rotors is displaced towards delivery 7 with respect to the centre of surface 810 of bearing 8.

For effective mounting between the two profiles of each rotor, bearing 8 may be made in two half-shells 81 , 82 which once coupled, for example by means of centring pins 70 and screws 71 , define a seat for said central part 2B, 2C of at least one of said rotors 2, 3.

Advantageously, as is seen in the figures, each half-shell 81 , 82 comprises two seats 8 A, 8B and 8C, 8D, each adapted to accommodate at least partially the central part 2C, 2D of the corresponding rotor. It should be noted that in figure 14, the representation of the bearing is simplified and a part of it is exploded only to improve the understanding of the drawing.

The hydrodynamic bearing described above reduces to almost zero the deflection of the rotors, allowing a very narrow clearance with a consequent increase in the volumetric efficiency of the pump and the possibility of working at pressures not obtainable with conventional pumps.

The hydrodynamic bearing is designed not only to support the load, but also to increase the rigidity of the single rotor during rotation, thus preventing dangerous phenomena and increasing the bending resistance, or the almost total insensitivity of any vibration phenomena to interference.

A particularly advantageous feature of the pump described herein is that the first 6A and the second 6B suction are mutually connected by a suction manifold 20 arranged outside the pump body 4 fed by a common suction 6C. This solution allows minimising the incoming turbulence of the fluid to be pumped and also allows implementing a structurally more rigid pump body, since the chambers and ducts necessary for suction and distribution of fluid intended to be processed by the rotors need not be provided inside the latter. This also ensures a compact pump body, leaving the suction part outside. In fact, in suction ducts, the speed of the incoming fluid is lower, and thus the sections must be larger and bulky. This leads to a great compactness of the body, resulting in increased mechanical strength (bending and torsional of the pump body), reduced vibrations, and any resonance frequencies all outside the operating range of the rotors. The suction capacity of the pump is increased, virtually excluding any problems of NPSH (Net Positive Suction Head) required in the operating range.

Figures 9 to 13 show s detail of the screw pumping profile of each rotor in detail. Advantageously, the pumping profiles of the rotors may be substantially equal and constructed so that the profiles of a rotor engage with those of the other, or profiles different from those described herein or only partially similar to them may be provided. Figure 9 shows an engaged part of the two rotors. Reference will be made only to two teeth 60 A and 60B, since the other ones of the profile(s) can be completely similar to them.

Specifically, teeth 60A, 60B seen in axial section are provided with a first concave side 63A, 63B and a second side 64A, 64B with a first convex part 64A1, 64B 1 and a second concave part 64A2, 64B2. Advantageously, the second concave part 64A2, 64B2 of the second side 64A, 64B is positioned in the proximity of a root 65 A, 65B of tooth 60A, 60B, while the first convex part 64A1, 64B 1 of the second side 64A, 64B is positioned in the proximity of a ridge 62A, 62B or apex of tooth 60A, 60B.

It is seen in the figures that the second concave part 64A2, 64B2 and the first convex part 64A1,

64B 1 may join at a central protruding area FA, FB of the tooth. In section, in fact, the central area is a point or portion of maximum lateral projection of the tooth. This central area is therefore a ridge on the side of the tooth.

With reference to figure 9, when the first 2 and the second 3 rotor are engaged, the first concave side 63A of tooth 60A of the first rotor may be facing the first concave side 63B of tooth 60B of the second rotor.

The first convex portion 64A1 of the second side 64 A of tooth 60 A of the first rotor 2 is facing the second concave part 64B2 of tooth 60B of the second rotor 3, while the second concave part 64A2 of tooth 60A of the first rotor is facing the first convex part 64B 1 of tooth 60B of the second rotor.

Advantageously, when the first and the second rotor are engaged, the distance between tooth 60A of the first rotor and tooth 60B of the second rotor is minimum, and thus a kind of seal (although not perfect) is formed in at least three points, which may be those enclosed in the circles in figure 9. This allows the pump to operate effectively on fluids also not very dense, such as for example water, which during the pumping remain trapped in volume P (fig. 9 and 11) between two engaged teeth.

Such zones or areas of minimum distance can be formed at the reciprocal central facing portions FA, FB (fig. 12) of the teeth, as well as at ridge 62A of tooth 60A of the first rotor facing root 65B of the tooth of the second rotor (fig. 11) and/or at the root of the tooth of the first rotor 65 A facing ridge 62B of the tooth of the second rotor (fig. 13).

In essence, according to the foregoing, the screw pumping profile of each rotor in axial section is provided with teeth 60A, 60B which, when mutually engaged, have at least three distinct points of minimum distance, intended to limit the leakage of fluid pumped during the pump operation. This allows great efficacy of the pumping effect, even in the presence of low density fluids, such as water.

As is seen, one of said points of minimum distance may be placed in the proximity of an apex of each tooth (fig. 11) and/or at a middle part of a side of each tooth (fig. 12) characterised by a ridge projecting on the side, and/or at a root of said tooth (fig. 13). Again with regard to the profile, it is noted that the basic equations are properly devised so that the two main escape routes of the fluid and therefore of reduction of the volumetric efficiency are closed (this happens to the outer diameter and the rolling diameter), within the strict limits of hundredth tolerances that still have to be provided, since the profiles should never touch during the rotation of the rotors. Advantageously, the processing and timing of these profiles - which are different for each side of the tooth - allows obtaining excellent pumping efficiency.

It should be noted that in section, the pump body has a rectangular or square shape. The "square" shape of the pump body leads to increase the inherent rigidity of the part containing the rotating members, with the consequent increase of reliability and performance by allowing higher rotational speeds than conventional pumps with oval or cylindrical section. Moreover, this shape increases the mechanical resistance both to internal pressures and to external loads (forces and moments), thereby allowing the pump to be able to operate with better performance. The support and fastening feet on the base are advantageously formed inside the pump body, thereby increasing the rigidity of the pump system. This further increases the resistance to external bending and torsional stresses, with undoubted gain for the reliability and the possibility of working in even more severe situations.

The bearings located in the heads are configured to ensure full support of the radial thrust and of any small axial thrust that may be temporarily generated due to the different trends of fluids at different viscosity in the operating spectrum. Advantageously, the bearings are all of the same type, symmetric, and with their ability they free the hydrodynamic bearing from any external action, thus increasing the functional performance thereof, and this also in function of the fact that the clearance of these bearings is lower, after assembly, than the radial one of the hydrodynamic bearing, which at this point is completely in charge only of the action of central support of the rotors and the consequent reduction of the camber. At the start of the pump, these bearings support all transient forces until the hydrodynamic bearing, having reached the operating speed, is capable of exerting all its bearing action.

According to an aspect of the invention, at least two teeth 60A, 60B intended to be mutually engaged, have each a first flat or concave side 63A, 63B and a second side 64A, 64B with a first convex or flat part 64A1, 64B1 and a second concave part 64A2, 64B2.

According to a further aspect, the first concave part 64A1, 64B 1 of the second side 64A, 64B is positioned in the proximity of a root 65 A, 65B of tooth 60A, 60B, while the second convex part 64A2, 64B2 of the second side 64A, 64B is positioned in the proximity of a ridge 62A, 62B of tooth 60A, 60B.

According to a further aspect, the first concave part 64A1, 64B1 and the second convex part 64A2, 64B2 join at a central zone FA, FB of the second side of tooth 64A.

According to yet another aspect, the first and the second rotor are engaged, the first concave side 63A of tooth 60A is facing the first concave side 63B of tooth 60B of the second rotor, and/or wherein when the first convex part 64 A 1 of the second side 64 A of tooth 60 A of the first rotor is facing the second concave part 64B2 of tooth 60B of the second rotor, and/or wherein the second concave part 64 A2 of tooth 60 A of the first rotor is facing the first convex part 64B 1 of tooth 60B of the second rotor.

According to a further aspect, the first and the second rotor are engaged, the distance between tooth 60A of the first rotor and tooth 60B of the second rotor is minimal at the reciprocal central facing portions FA, FB, and/or at ridge 62A of tooth 60A of the first rotor facing root 65B of the second rotor, and/or at the root of the tooth of the first rotor 65A facing ridge 62B of the tooth of the second rotor. Various embodiments of the invention have been described but others may be conceived using the same innovative concept.

Claims

Positive displacement pump (1) comprising at least a first rotor (2) and a second rotor (3) at least partially sealingly accommodated inside a pump body (4), the first and second rotor having synchronization means (5) of their rotation, each rotor comprising a first (2A, 3A) and a second (2B, 3B) screw pumping profile spaced by a central part (2C, 3C) without pumping profile, the first profile (2A) and the second profile (2B) of the first rotor

(2) being engaged with the first profile (3A) and the second profile (3B) of the second rotor (3), respectively, so as to pump a fluid processable by the pump by sucking it from a first (6A) and a second (6B) suction associated to the end sides of the rotors (2,

3) and pumping it towards said central part (2C, 3C) to which a delivery (7) of the pump is associated, said central part (2C, 3C) of each rotor being coupled to a hydrodynamic bearing (8)accommodated inside the pump body (4), characterized in that said hydrodynamic bearing is operated by a part of the fluid processed by the pump, wedged between the bearing (8) and the central part (2B, 2C) of the rotors.

Pump according to the preceding claim, wherein the hydrodynamic bearing (8) comprises at least one inner cylindrical wall (810), intended to hydrodynamically cooperate with said central part (2B, 2C) to support it and an outer cylindrical perimeter wall (811) which can be associated to the pump body (4).

Pump according to the preceding claim, wherein at each of the ends of the bearing facing each pumping profile, at least one chamber (10A, 10B) is provided, intended to accommodate at least a part of the fluid under pressure, the chamber (10A, 10B) being in communication with said delivery (7).

4. Pump according to claim 2, wherein said inner cylindrical wall (810) is eccentric with respect to said central part of the rotor and has a greater diameter than that of said central part (2B, 2C).

5. Positive displacement pump according to claim 1, wherein the hydrodynamic bearing (8) consists of at least two half-shells (81 , 82) which when coupled define a seat for said central part (2B, 2C) of at least one of said rotors (2, 3).

6. Pump according to the preceding claim, wherein each half-shell (81, 82) comprises two seats (8A, 8B, 8C, 8D), each adapted to accommodate at least part of the central part (2C, 2D) of the corresponding rotor.

7. Pump according to one or more of the preceding claims, wherein the mutual synchronization means (5) of rotation comprise herringbone gear, preferably stopped by locking assemblies with automatic compensation of the axial thrust.

8. Pump according to one or more of the preceding claims, wherein the first (6 A) and the second (6B) suction are mutually connected by a suction manifold (20) arranged outside the pump body (4) fed by a common suction (6C), and/or wherein the pump body (4) has a square or rectangular sectional shape.

9. Pump according to one or more of the preceding claims, wherein the screw pumping profile of each rotor in axial section is provided with teeth (60A, 60B) which, when mutually engaged, define at least three distinct areas of minimum distance, intended to limit the leakage of fluid pumped during the pump operation.

10. Pump according to the preceding claim, wherein at least one of said areas of minimum distance is placed in the proximity of an apex of each tooth, and/or at a middle part of a side of each tooth, and/or at a root of said tooth.