ES2764114T3 - Centrifugal pump to transport a highly viscous fluid - Google Patents

Abstract

A centrifugal pump for conveying a highly viscous fluid having a kinematic viscosity of at least 10-4 m2 / s, including a casing (2) with at least a first inlet (3) and an outlet (4) for the fluid, a impeller (5) for conveying fluid from inlet (3) to outlet (4), where impeller (5) is arranged on a rotary shaft (6) for rotation about an axial direction (A), and includes a front casing (7) facing the first inlet (3) of the pump, where the casing (2) is provided with a stationary impeller opening (8) that receives the front casing (7) of the impeller (5) and having a diameter (D), where the front casing (7) and the stationary impeller opening (8) form an interval (9) having a length (L) in the axial direction (A), characterized in that the interval (9) extends parallel to the axis (6), and because the ratio of the length (L) of the gap (9) and the diameter (D) of the impeller opening (8) is at or I add 0.092.

Description

DESCRIPTION

Centrifugal pump to transport a highly viscous fluid

The invention relates to a centrifugal pump for conveying a highly viscous fluid according to the preamble of the independent claim.

Pumps for pumping highly viscous fluids are used in many different industries, for example in the oil and gas processing industry to transport hydrocarbon fluids. Here, these pumps are used for different applications such as extracting crude from an oil field, transporting oil or other hydrocarbon fluids through pipelines or within refineries. But also in other industries, for example the food industry or the chemical industry, highly viscous fluids often have to be transported. An example of such a pump is described in DE4208202, which describes a centrifugal pump with an annular seal between a rotor and a casing. The annular seal has a clearance at least adjacent to a rotor inlet that is tilted, in the direction of a main flow of fluid at the rotor inlet, at an angle between 0 degrees and 60 degrees relative to the axis of rotation of the rotor.

Another example of such a pump is described in CN101892989, which describes a high pressure, double suction pump. The pump includes a suction chamber, a pump casing, an impeller, a shaft, a packaged mechanical sealing device arranged between the pump casing and the shaft, a impeller ring sealing device, and a sealing flushing device. mechanical, where the suction chamber includes a suction port, a flow distribution section and a spiral section; the flow division section includes a left pass, a right pass, and a volute wall between the left pass and the right pass; the volute wall is provided with a water distribution tab that extends into the suction port; and the outer surface of the flow distribution section of the suction chamber is provided with at least one radial reinforcing rib that surrounds the flow distribution section with a circle.

Fluid viscosity is a measure of internal friction generated in a circulating fluid and a characteristic property of the fluid. Within the structure of this application, the terms "viscosity" or "viscous" are used to designate the kinematic viscosity of the fluid and the term "highly viscous fluid" will be understood as meaning that the fluid has a kinematic viscosity of at least 10-4 m2 / s, which is 100 centistokes (cSt).

The use of centrifugal pumps for the pumping of highly viscous fluids is known. Pumping highly viscous fluids with centrifugal pumps requires a considerably higher power pump than, for example, pumping water. The higher the viscosity of the fluid, the greater the power required by the pump to supply the required pumping volume. Especially in the oil and gas industry, the main focus, at least in the past, has been on the volume of pumping, i.e. the flow generated by the pump, and on the reliability of the pump rather than on the pump efficiency. However, nowadays an attempt is made to achieve a more efficient use of the pump. It is desirable to have the highest possible ratio of power, especially hydraulic power, delivered by the pump to the power needed to move the pump. This desire is mainly based on a greater awareness of environmental protection and a responsible use of available resources, as well as the increasing costs of energy.

To improve the efficiency of a centrifugal pump for pumping highly viscous fluids it is known to use specific impeller designs, especially impellers with high load coefficients. The load coefficient of the impeller can be increased, for example, by increasing the exit angle of the vane or the number of vanes or the exit width of the impeller. Despite these measures, the efficiency of a pump for pumping highly viscous fluids still needs to be further improved.

Therefore, an object of the invention is to propose a new centrifugal pump for transporting highly viscous fluids that has a better efficiency, that is, an increased ratio of the power distributed by the pump when pumping the fluid to the power that is supplied to the pump to move the pump

The subject matter of the invention which achieves this object is characterized by the elements of the independent claim. Thus, according to the invention, a centrifugal pump is proposed to transport a highly viscous fluid that has a kinematic viscosity of at least 10-4 m2 / s, including a housing with at least a first inlet and an outlet for the fluid, an impeller for transporting the fluid from the inlet to the outlet, where the impeller is arranged on a rotating shaft for rotation about an axial direction, and includes a front casing facing the first inlet of the pump, where the casing is provided with an opening stationary impeller to receive the front impeller casing and having a diameter, where the front casing and the stationary impeller opening form an interval having a length in the axial direction, characterized in that the interval extends parallel to the axis, and in that the ratio of the interval length and the diameter of the impeller opening is at least 0.092.

The invention is based in particular on the finding that pump efficiency can be increased by pumping highly viscous fluids by designing the gap between the impeller front casing and the stationary impeller opening considerably shorter than has been done in the art previous.

The interval, which is also sometimes called the labyrinth, is necessary to seal the high pressure side of the impeller, more specifically the side space, against the pump inlet. The impeller is disposed in the stationary impeller opening which is a part of the pump which is stationary with respect to the casing and which is adapted to receive the impeller. In the assembled state, the impeller is located in said impeller opening such that the gap or labyrinth is between the outer circumferential surface of the forward impeller shell and the inner circumferential surface of the stationary impeller opening. This interval has a length in the axial direction that seals between the side space on the high pressure side of the impeller and the pump inlet, which is the low pressure side of the pump.

During pump operation, backflow is generated from the high pressure side of the impeller, which, in a single stage pump, is the region near the pump outlet, through the lateral space, and through of the gap between the front casing and the stationary impeller opening, returning to the low pressure side of the impeller.

The gap or labyrinth, respectively, is designed as a radial clearance gasket or labyrinth, that is, it provides clearance with respect to the radial direction. Therefore, the main flow through the interval is in the axial direction, that is, parallel to the axis. This has to be differentiated from an axial clearance joint or labyrinth that extends perpendicular or obliquely to the axis, so that the main flow through an axial clearance joint is in a radial or oblique direction with respect to the radial direction. In an axial clearance joint, the clearance in the axial direction changes with relative movement of the stationary part and the turning part in the axial direction, where in a radial clearance joint the clearance in the radial direction changes with the relative movement of the stationary part and the turning part in the radial direction.

An essential finding is that, due to the short axial length of the interval (i.e. the labyrinth) proposed by the invention, the power losses through the interval decrease, among others, due to the reduced drag resistance in the lateral space. On the other hand, shortening the interval can be expected to result in a reduced sealing action, thus increasing the reflux in the pump. However, an increased reflux flow rate reduces pump efficiency and therefore runs counter to improved efficiency. Therefore, the unexpected finding is that, by shortening the interval with respect to the axial direction, the overall pumping efficiency increases despite the risk of increased reflux flow.

According to the invention, the length of the gap should not exceed 0.092 times the diameter of the impeller opening.

The optimal length of the interval depends on several factors, for example, the viscosity of the fluid. Thus, depending on the specific application, it may be preferable that the ratio of the length of the gap and the diameter of the impeller opening is at most 0.073 and preferably at most 0.055.

There are also applications for which it is advantageous that the ratio of the length of the gap and the diameter of the impeller opening is at most 0.037 and preferably at most 0.019.

For practical reasons there is also a preferred lower limit for the length of the interval. According to the preferred design, the ratio of the interval length and the diameter of the impeller opening is at least 0.0001.

In order to generate the desired sealing effect for the interval it is preferred to have a radial clearance between the front shell and the impeller opening that is at most 0.0045 times the diameter of the impeller opening. Radial clearance is the extension of the interval with respect to the radial direction, that is, perpendicular to the axial direction, and can be considered as the width of the interval. This radial clearance is the minimum distance between the outer circumferential surface of the forward impeller shell and the inner circumferential surface of the stationary impeller opening over the interval.

The two surfaces that delimit the interval can be designed as level surfaces.

According to another embodiment, the interval includes a plurality of flat parts consecutively arranged with respect to the axial direction, where two adjacent flat parts are respectively separated by a slot. In such an embodiment, the two surfaces delimiting the interval are not level. The part of the outer circumferential surface of the front part of the impeller that delimits the interval or the part of the internal circumferential surface of the stationary opening of the impeller that delimits the interval can be provided with a plurality of flat parts and grooves in between. In such an embodiment the length of the gap in the axial direction is defined as the sum of the lengths of all the individual flat parts in the axial direction. The grooves do not contribute to the overall length of the gap in the axial direction.

According to a preferred embodiment, the stationary inlet opening includes a wear ring that delimits the interval with respect to the radial direction, the wear ring being arranged stationary with respect to the housing.

In addition or as an alternative measure, it is also possible that the impeller includes a wear ring that delimits the interval with respect to the radial direction, the wear ring being arranged stationary with respect to the impeller.

The centrifugal pump can be designed, for example, as a single suction pump or a double suction pump, as a single stage pump or as a multistage pump. When the pump is designed as a single suction pump, it can have a rear casing on the impeller in addition to the front casing. In such a design it is also possible for the rear impeller shell to form a gap with a part that is stationary with respect to the housing. This interval in the rear casing can be designed in the same analogous way, as explained with respect to the interval in the front casing of the impeller.

According to a preferred embodiment, the centrifugal pump is designed as a double suction pump, having a second fluid inlet that is arranged opposite to the first pump inlet, where the impeller is designed as a double suction impeller. including vanes to transport the fluid both from the first inlet and from the second inlet to the outlet.

For such a design as a double suction pump it is preferred that the impeller includes a second front casing oriented to the second inlet of the pump, where the casing is provided with a second stationary impeller opening to receive the second front casing of the impeller and which has a diameter, where the second front casing and the second stationary impeller opening form a second interval having a length in the axial direction, and where the ratio of the length of the second interval and the diameter of the second impeller opening is a at most 0.092.

Depending on the specific application, it may be preferable that also the ratio of the length of the second gap and the diameter of the second impeller opening is at most 0.073 and preferably at most 0.055. There are also applications for which it is advantageous that the ratio of the length of the second gap and the diameter of the second impeller opening is at most 0.037 and preferably at most 0.019.

Also for the second interval it is advantageous that there is a radial clearance between the second front casing and the second impeller opening that is at most 0.0045 times the diameter of the second impeller opening. An especially preferred measure is that the interval and the second interval are designed essentially identically.

According to an essential application, the pump is designed for use in the oil and gas industry.

Other advantageous measures and embodiments of the invention will be apparent from the dependent claims. The invention will be explained in more detail below with reference to the drawings. In it they are shown in a schematic representation:

Fig. 1 is a cross sectional view of an embodiment of a pump according to the invention.

Figure 2 is an enlarged representation of the detail I of figure 1.

Figure 3 is an outline of the front shell and a wear ring as part of the stationary impeller opening.

Figure 4 is like Figure 3, but for a variant of the embodiment.

Figure 5 is a second variant for the design of the gap between the front casing and the stationary impeller opening.

And Figure 6 is an illustration of a comparison of a pump according to the invention with prior art pumps.

Figure 1 represents a cross-sectional view of an embodiment of a centrifugal pump according to the invention indicated in its entity with the reference number 1. Figure 2 shows an enlarged representation of the detail I of Figure 1. Pump 1 It is designed to carry a highly viscous fluid, while the term "highly viscous" means that the kinematic viscosity of the fluid is at least 10-4 m2 / s, which is 100 centistokes (cSt).

In this embodiment, pump 1 is designed as a single stage, double suction centrifugal pump. This design is a preferred embodiment that is useful in practice for many applications. Naturally, the invention is not limited to this design. A pump according to the invention can also be designed as a single suction centrifugal pump or as a multistage centrifugal pump or as any other type of centrifugal pump. Based on the description of the embodiment represented in figure 1 and figure 2, it is not a problem for experts to devise a pump according to the invention, which is designed as another type of pump, especially centrifugal pump, for example, a suction pump only.

The double suction pump 1 includes a casing 2 with a first inlet 3, a second inlet 3 'and an outlet 4 for the fluid to be pumped. The fluid can be, for example, crude oil, oil, or any other highly viscous hydrocarbon fluid. The pump 1 has an impeller 5 with a plurality of vanes 51 to transport the fluid from the first inlet 3 and the second inlet 3 'to the outlet 4. The impeller 5 is arranged on a rotating shaft 6 for rotation about an axial direction A. Axial direction A is defined by the axis of axis 6 about which impeller 5 rotates during operation. Axis 6 is rotated by a drive unit (not shown). The direction perpendicular to the axial direction A is called the radial direction.

The first inlet 3 and the second inlet 3 'are arranged opposite each other with respect to the axial direction A. Thus, according to the representation of figure 1, the fluid flows both from the left and from the right side in the axial direction A to the impeller 5, while the fluid from the first inlet 3 flows in the opposite direction to the impeller 5 as the fluid from the second inlet 3 '. The impeller 5 transports the fluid from the first inlet 3 and the fluid from the second inlet 3 'in the radial direction to the outlet 4 of the pump.

The impeller 5 includes a front casing 7 covering the vanes 51 and oriented to the first inlet 3 of the pump 1. Since in this embodiment the impeller 5 is designed as a double suction impeller 5 it includes a second front casing 7 'oriented to the second inlet 3 'and covering the vanes 51 on the side of the impeller 5 facing the second inlet 3'.

The casing 2 is provided with a stationary impeller opening 8 to receive the front casing 7 of the impeller 5. The stationary impeller opening 8 is stationary with respect to the casing 2 of the pump 1 and has a circular cross section with a diameter D , while the diameter D designates the smallest diameter of the part of the stationary impeller opening 8 that receives the front casing 7.

Similarly, housing 2 includes a second stationary impeller opening 8 'to receive the second front casing 7' of impeller 5.

In the assembled state, the impeller 5 is arranged coaxially within the stationary impeller opening 8 such that the outer circumferential surface of the front casing 7 faces the inner circumferential surface of the stationary impeller opening 8. Thus, the casing front 7 and stationary impeller opening 8 form an interval 9 (see also FIG. 3) between front casing 7 and stationary impeller opening 8. Interval 9 is also called a labyrinth. It is essentially annular in shape and performs a sealing action, as will be explained below. Interval 9 has a length L which is the extension of interval 9 in axial direction A. Interval 9 extends parallel to axis 6 or parallel to axial direction A, respectively. Thus, the reflux flows through the interval 9 parallel to axis 6 and in the opposite direction when the fluid flows through the respective inlet 3. Thus, seen in the main direction of flow of the fluid entering through the respective inlet 3, the initial position of interval 9, i.e. the opening through which the fluid enters interval 9, is behind the final position of interval 9, i.e. the opening through which the fluid exits of interval 9.

Similarly, a second gap 9 'is formed between the second front casing 7' and the second stationary impeller opening 8 '. The second interval 9 'has a length L' in the axial direction A and the second stationary impeller opening 8 'has a diameter D'. The interval 9 'extends parallel to axis 6 or parallel to axial direction A, respectively. Preferably, but not necessarily, length L 'is equal to length L and diameter D' is equal to diameter D. Since the design and arrangement of the second interval 9 'may be identical to interval 9, the following description only it will refer to interval 9. This description will be understood to apply analogously also to the second interval 9 '.

The gap 9 or labyrinth 9 seals a side space 10 located on the high pressure side of the impeller 5 against the low pressure side of the impeller 5 which is located at the inlet 3. The side space 10 is located on the high side pressure of the impeller 5 near the outlet 4 of the pump 1 and delimited by the front casing 7 of the impeller 5 as well as by the casing 2 of the pump 1. During the operation of the pump 1, a reflux is generated from the area of the outlet 4 through the lateral space 10. The reflux passes through the interval or the labyrinth 9 flowing essentially in the axial direction A, that is, parallel to the axis 6 and arrives at the low pressure side of the impeller 5 next to the first inlet 3. It is obvious that reflux reduces the efficiency of pump 1.

Thus, one of the functions of interval 9 is to provide some sealing action to limit backflow. That is the reason why interval 9 is also called a maze.

The basic idea of the present invention is to shorten the lengths L (see figure 2 and figure 3) of the interval 9 in the axial direction A compared to the known solutions of the prior art. Although it would be expected that a shortening of the length L would lead to increased reflux which, in turn, reduces the pumping efficiency, it has been observed that, by shortening the length L of the interval 9, the overall efficiency of the pump 1 can be increased.

With reference to Figure 2 and Figure 3, the design of the interval 9 will now be explained in more detail. In the embodiment according to figure 1, the stationary inlet opening 8 includes a wear ring 11 that delimits the interval 9 with respect to the radial direction. The wear ring 11 faces the outer circumferential surface of the front casing 7 which is inserted into the stationary inlet opening 8. The wear ring 11 is fixedly mounted on the housing 2, thus, the wear ring 11 is stationary with relative to housing 2.

Figure 3 depicts an outline of the front shell 7 and wear ring 11 as part of the stationary impeller opening 8 to better understand the dimensions of the gap 9.

According to the invention, the length L of the interval 9 is designed in such a way that the ratio of the length L and the diameter D of the impeller opening 8 is at most 0.092, that is, L / D <0.092. As already indicated, the diameter D designates the smallest diameter of the stationary impeller opening 8, that is, the diameter in the position where the wear ring 11 is closest to the outer circumferential surface of the front casing 7. The length L of the interval 9 is the extension in the axial direction A of the zone where the stationary impeller opening 8 and the front casing 7 are closest to each other.

In the arrangement shown in Figure 3, the wear ring 11 is designed with a projection 111 in the radial direction. Accordingly, the length L of the interval 9 is equal to the extension of the projection 111 in the axial direction 9. The second parameter defining the geometry of the interval 9 is the radial clearance R between the front casing 7 and the stationary impeller opening 8 or the wear ring 11, respectively, along the axial extension of the interval 9. The radial clearance R designates the minimum radial clearance along the interval 9.

It has been shown in practice that it is advantageous if the radial clearance R does not exceed 0.0045 times the diameter D of the stationary inlet opening 8, ie the condition R / D <0.0045 is preferably met.

The optimal length L of the interval 9 depends on the respective application. There are several factors that influence an appropriate choice of the length L of the interval 9, for example, the kinematic viscosity of the specific fluid to be pumped, the pressure increase generated by the pump, the flow through the pump or other operating parameters of pump 1.

For a given set of pump 1 operating parameters, the lengths L of range 9 should preferably be reduced with increasing viscosity of the fluid to be pumped.

In practice and depending on the application it may be preferable that the L / D ratio does not exceed 0.073 or more preferably that it does not exceed 0.055, or even more preferably that it does not exceed 0.037 or it is specifically preferred that it does not exceed 0.019.

According to the preferred embodiments of pump 1, the minimum L / D ratio is 0.0001, that is, the length L of interval 9 is preferably at least 0.0001 times the diameter of the stationary impeller opening 8 or the ring of wear 11, respectively.

Figure 4 shows in a representation similar to Figure 3 a variant of the embodiment of the pump 1. According to this variant, the impeller 5 and more specifically the front casing 7 of the impeller 5 includes a wear ring 11 'that delimits the interval 9 with respect to the radial direction. The wear ring 11 'is fixedly connected to the impeller 5 and rotates with the impeller 5. In this variant, the stationary impeller opening 8 can also include a wear ring 11, but can also be designed without a wear ring.

Figure 5 illustrates a second variant for the design of the gap 9 between the front casing 7 and the stationary impeller opening 8. According to the second variant, the stationary impeller opening 8 or the wear ring 11, respectively, or as an alternative (not shown) the front shell 7 is designed in such a way that the interval 9 includes a plurality of flat parts12 arranged consecutively with respect to the axial direction A, where two adjacent flat parts 12 are respectively separated by a slot 13. In such a design , the total length L of the interval 9 is the sum of the individual lengths L1, L2, L3, L4, L5 of all the flat parts 12 in the axial direction. The extension of the grooves does not contribute to the total lengths L of interval 9, that is, L = L1 + L2 + L3 + L4 + L5. It will be understood that the number of flat parts and grooves, as well as their geometric design represented in figure 5, is only exemplary.

The pump 1 according to the invention has a better pumping efficiency compared to the pumps known in the current art. Pumping efficiency designates the ratio of the power delivered by the pump to the power input for the pump, that is, the power used to drive the pump. The power distributed by the pump is generally the hydraulic power generated by pump 1.

Figure 6 illustrates a comparison of a pump according to the invention with pumps of the prior art. The graph represents the pumping efficiency P as a function of the viscosity V of the fluid carried by the pump. For a better understanding, the graph is standardized in such a way that the pumping efficiency P of the prior art pumps is equal to the horizontal viscosity axis V, that is, the pumping efficiency P for the pump according to the prior art is always on the V axis for each viscosity. Thus, the graph directly represents the increase in the pumping efficiency of pump 1 according to the invention compared to a prior art pump. The pumping efficiency of the pump according to the invention is represented by curve K. As can be clearly seen, as soon as the fluid viscosity is greater than a specific value V1, the pump 1 according to the invention has a higher pumping efficiency. compared to the prior art pump. The efficiency gain increases with the viscosity of the fluid. The specific value V1 of the viscosity where the pump 1 according to the invention is more efficient than the prior art pump is generally less than the value of 10-4 m2 / s. Thus, for a highly viscous fluid, the pump 1 according to the invention has a higher pumping efficiency than the prior art pump.

Although specific reference has been made for purposes of explanation to an embodiment, where pump 1 is designed as a single stage, double suction centrifugal pump, the invention is in no way limited to such embodiments. The pump according to the invention can also be designed like any other type of centrifugal pump, for example, as a single suction pump or as a multistage pump. In particular, the invention is applicable both to centrifugal pumps with a closed impeller, i.e., an impeller having a front casing and a rear casing, and to centrifugal pumps with a semi-open impeller, i.e., having a rear casing, but not wrapped front. In such designs where the impeller has a rear shell or a rear shell only, the gap design 9 according to the invention can be used for the rear shell in a manner analogous to that described herein with reference to the front shell.

Claims (15)

1. A centrifugal pump to transport a highly viscous fluid that has a kinematic viscosity of at least 10-4 m2 / s, including a casing (2) with at least a first inlet (3) and an outlet (4) for the fluid , an impeller (5) to transport the fluid from the inlet (3) to the outlet (4), where the impeller (5) is arranged on a rotary axis (6) for rotation about an axial direction (A), and it includes a front casing (7) oriented to the first inlet (3) of the pump, where the casing (2) is provided with a stationary impeller opening (8) that receives the front casing (7) of the impeller (5) and having a diameter (D), where the front casing (7) and the stationary impeller opening (8) form an interval (9) that has a length (L) in the axial direction (A), characterized in that the interval ( 9) extends parallel to the axis (6), and because the ratio of the length (L) of the interval (9) and the diameter (D) of the impeller opening (8) is at most 0.092. 2. The centrifugal pump according to claim 1, wherein the ratio of the length (L) of the interval (9) and the diameter (D) of the impeller opening (8) is at most 0.073 and preferably at most 0.055. The centrifugal pump according to any of the preceding claims, wherein the ratio of the length (L) of the interval (9) and the diameter (D) of the impeller opening (8) is at most 0.037 and preferably at most 0.019. The centrifugal pump according to any of the preceding claims, wherein the ratio of the length (L) of the interval (9) and the diameter (D) of the impeller opening (8) is at least 0.0001. The centrifugal pump according to any of the preceding claims, having a radial clearance (R) between the front casing (7) and the impeller opening (8) which is at most 0.0045 times the diameter (D) of the impeller opening (8). The centrifugal pump according to any of the preceding claims, wherein the interval (9) includes a plurality of flat parts (12) consecutively arranged with respect to the axial direction (A) and where two adjacent flat parts (12) are respectively separated through a slot (13). The centrifugal pump according to any of the preceding claims, wherein the stationary inlet opening (8) includes a wear ring (11) that delimits the interval (9) with respect to the radial direction, the wear ring (11) being arranged stationary with respect to the casing (2). The centrifugal pump according to any of the preceding claims, wherein the impeller (5) includes a wear ring (11 ') that delimits the interval (9) with respect to the radial direction, the wear ring (11') being arranged ) stationary with respect to the impeller (5). The centrifugal pump according to any of the preceding claims, which is designed as a double suction pump, having a second inlet (3 ') for the fluid that is arranged opposite to the first inlet (3) of the pump, where the impeller (5) is designed as a double suction impeller (5) including vanes (51) to transport the fluid from both the first inlet (3) and the second inlet (3 ') to the outlet (4). The centrifugal pump according to claim 9, where the impeller (5) includes a second front casing (7 ') oriented to the second inlet (3') of the pump, where the casing (2) is provided with a second opening stationary impeller (8) to receive the second front casing (7 ') of the impeller and having a diameter (D'), where the second front casing (7 ') and the second stationary impeller opening (8') form a second interval (9 ') having a length (L') in the axial direction (A), and where the ratio of the length (L ') of the second interval (9') and the diameter (D ') of the second impeller opening (8 ') is at most 0.092. 11. The centrifugal pump according to claim 9 or 10, wherein the ratio of the length (L ') of the second interval (9') and the diameter of the second impeller opening (8 ') is at most 0.073 and preferably at at most 0.055. 12. The centrifugal pump according to any of claims 9-11, where the ratio of the length (L ') of the second interval (9') and the diameter (D ') of the second impeller opening (8') is a at most 0.037 and preferably at most 0.019. 13. The centrifugal pump according to any of claims 9-12 having a radial clearance between the second front casing (7 ') and the second impeller opening (8') which is at most 0.0045 times the diameter (D ') of the second impeller opening (8). 14. The centrifugal pump according to any of claims 9-12, wherein the interval (9) and the second interval (9 ') are essentially identical in design. 15. The centrifugal pump according to any of the preceding claims which is designed for use in the oil and gas industry.