Patent Publication Number: US-2022220980-A1

Title: Rotary pump for conveying a fluid

Description:
CROSS-REFERENCE TO RELATED APPLICATION 
     This application claims priority to European Patent Application No. 21151344.5, filed Jan. 13, 2021, the contents of which are hereby incorporated by reference in their entirety. 
     BACKGROUND 
     Field of the Invention 
     The disclosure relates to a rotary pump for conveying a fluid. 
     Background Information 
     Rotary pumps for conveying a fluid, for example a liquid such as water, are used in many different industries. Examples are the oil and gas industry, the power generation industry, the chemical industry, the water industry or the pulp and paper industry. Conventional rotary pumps have at least one impeller and a pump shaft for rotating the impeller. The at least one impeller can be configured for example as a radial impeller or as an axial or semi-axial impeller or as a helicoaxial impeller. Furthermore, the impeller can be configured as an open impeller or as a closed impeller, where a shroud is provided on the impeller, said shroud at least partially covering the vanes of the impeller. 
     A rotary pump can be designed as a single stage pump having only one impeller mounted to the shaft or as a multistage pump comprising a plurality of impellers, wherein the impellers are arranged one after another on the shaft. The impellers can be arranged in an in-line arrangement, where the axial thrust generated by a single impeller is directed in the same direction for all impellers, or in a back-to-back arrangement, where the axial thrust generated by a first group of impellers is directed in the opposite direction as the axial thrust generated by a second group of impellers. 
     Many rotary pumps are provided with at least one balancing device or balancing system for at least partially balancing the axial thrust that is generated by the impeller(s) during operation of the pump. The balancing device reduces the axial thrust that has to be carried by the axial bearing or the thrust bearing. The balancing device can comprise a balance drum for at least partially balancing the axial thrust that is generated by the rotating impellers. The balance drum is fixedly connected to the pump shaft of the pump in a torque proof manner. Usually, in a single stage pump or in a multistage pump with in-line arrangement of the impellers the balance drum is arranged at the discharge side of the pump between the last stage impeller and a shaft sealing device. In a multistage pump with a back-to-back arrangement of impellers the balance drum is usually located adjacent to an intermediate stage impeller, which is arranged at one end of the hydraulic unit comprising all the impellers. The balance drum defines a front side and a back side. The front side is the side facing the hydraulic unit. The back side is the side facing the shaft sealing device. 
     A relief passage is provided between the balance drum and a stationary part being stationary with respect to the pump housing. The back side is usually connected to the suction side or a low pressure location of the pump by means of a balance line. At the low pressure location a pressure prevails, which is smaller than the pressure at the front side. During operation there is a leakage flow through the relief passage from the front side along the balance drum to the back side and from there through the balance line to the suction side. At the front side of the balance drum the higher pressure or the discharge pressure prevails, and at the back side essentially the suction pressure or the low pressure prevails. The pressure difference between the front side and the back side results in an axial force or an axial thrust which is directed in the opposite direction as the axial thrust generated by the rotating impeller(s). Thus, the axial thrust that has to be carried by the axial or thrust bearing is at least considerably reduced. Of course, the leakage flow along the balance drum results in a decrease of the hydraulic performance or efficiency of the pump. Therefore, the relief passage is configured such, that the leakage flow is low but still sufficient for generating the axial thrust counteracting the axial thrust generated by the impeller(s). 
     SUMMARY 
     It has been determined that nowadays in many applications the most efficient use of the pump is strived for. It is desirable to have the highest possible ratio of the power, especially the hydraulic power, delivered by the pump to the power needed for driving the pump. This desire is mainly based upon an increased awareness of environment protection and a responsible dealing with the available resources as well as on the increasing costs of energy. As already said, the flow of the fluid passing along the balance drum through the relief passage, which is in most cases the main leakage flow occurring in the pump, reduces the efficiency of the pump. 
     It is therefore an object of the disclosure to propose a rotary pump for conveying a fluid, having a reduced leakage flow through the balancing system and therewith an increased efficiency without reducing the balancing of the axial thrust acting on the pump shaft during operation of the pump. 
     The subject matter of embodiments of the invention satisfying this object is characterized by the features disclosed herein. 
     Thus, according to an embodiment of the invention, a rotary pump for conveying a fluid is proposed, comprising a pump housing with an inlet for receiving the fluid having a suction pressure, an outlet for discharging the fluid having a discharge pressure, a pump shaft configured for rotating about an axial direction, and a hydraulic unit for conveying and pressurizing the fluid, wherein the hydraulic unit comprises at least one impeller fixedly mounted on the pump shaft, the pump further comprising a balance drum fixedly connected to the pump shaft and arranged between the hydraulic unit and an end of the pump shaft, wherein the balance drum defines a front side facing the hydraulic unit and a back side facing away from the hydraulic unit, wherein an axial relief passage is diposed between the balance drum and a stationary part configured to be stationary with respect to the pump housing, wherein a balance line is provided connecting the back side with a low pressure location, wherein an additional balancing device is arranged between the balance drum and the hydraulic unit, the additional balancing device comprising a ring-shaped rotary part fixedly connected to the pump shaft, and a ring-shaped non-rotary part, which is movable only in the axial direction, wherein the rotary part has an axial face facing the hydraulic unit, wherein the rotary part and the non-rotary part are configured to overlap with respect to a radial direction, which is perpendicular to the axial direction, and wherein the non-rotary part is configured to be movable in the axial direction such that a radial relief passage is open during operation of the pump, with the radial relief passage extending in the radial direction between the rotary part and the non-rotary part. 
     The combined balancing system comprising the balance drum and the additional balancing device is located adjacent to the hydraulic unit with the at least one impeller. The non-rotary part of the additional balancing device is configured such that it is movable forth and back in the axial direction but cannot rotate. Depending of the specific embodiment of the pump the movement of the non-rotary part can be caused for example by hydraulic forces only, or by a combination of at least two different types of force, e.g. a hydraulic force in combination with a spring force, or—in particular in a vertical pump with the pump shaft extending in the vertical direction (direction of gravity)—a hydraulic force in combination with a gravitational force, or a hydraulic force in combination with a magnetic force. Of course, also other combinations of forces can be used to move the non-rotary part. According to a preferred embodiment, the non-rotary part is spring-loaded by a spring element, wherein the spring element exerts a force on the non-rotary part, which is directed towards the rotating part of the additional balancing device. Preferably the spring element is configured such that the non-rotary part is pressed against the rotary part of the additional balancing device at standstill of the pump. 
     During start-up of the pump the hydraulic force acting both on the non-rotary part and the rotary part increases until the pump reaches its nominal speed. Since the non-rotary part of the additional balancing device is movable in the axial direction the hydraulic force pushes the non-rotary part away from the rotary part for example against the force of a spring element or against the force of a magnet, such as a permanent magnet, so that the radial relief passage opens between the rotary part and the non-rotary part of the additional balancing device. The fluid flows through the radial relief passage, which is also referred to as radial labyrinth, to the front side of the balance drum and then through the axial relief passage along the balance drum to the back side and into the balance line. 
     By this combination of an radial relief passage and an axial relief passage, wherein the radial relief passage is arranged between the hydraulic unit and the axial relief passage the overall leakage flow through the additional balancing device and along the balance drum can be considerably reduced as compared to known balancing systems having a balance drum only. Although the overall leakage flow is considerably reduced the balancing action regarding the axial thrust is at least not reduced as compared to known balancing systems. 
     Within this application a “radial relief passage” or “radial gap” or a “radial labyrinth” designates a passage which extends in the radial direction, such that the fluid passing through said passage flows in radial direction, i.e. in a direction perpendicular to the pump shaft. 
     Furthermore, within this application an “axial relief passage” or an “axial gap” or an “axial labyrinth” designates a passage which extends in the axial direction, such that the fluid passing through said passage flows in the axial direction, i.e. in a direction parallel to the pump shaft. 
     According to a preferred configuration, the balance drum delimits an annular chamber arranged at the front side, wherein the annular chamber extends between the rotary part and the non-rotary part with respect to the radial direction. Thus, during operation of the pump, in the annular chamber an intermediate pressure prevails, which is smaller than the pressure acting on the axial face of the rotary part, e.g. the discharge pressure, and which is larger than the pressure at the low pressure location, e.g. the suction pressure. 
     Preferably, the non-rotary part comprises a first axial face and a second axial face delimiting the non-rotary part with respect to the axial direction, wherein the first axial face is arranged to be exposed to the same pressure as the axial face of the rotary part facing the hydraulic unit, and wherein the second axial face is arranged to be exposed to the pressure prevailing in the annular chamber, i.e. the intermediate pressure. Thus, the width of the radial relief passage, i.e. the extension of the radial relief passage in the axial direction is self-adjusting. 
     If, during operation of the pump, the pressure acting of the first axial face of the non-rotary part increases, the width of the radial relief passage increases, meaning that the radial relief passage becomes broader with respect to the axial direction. Consequently the leakage flow through the radial relief passage increases, whereby the resistance for the fluid flowing through the axial relief passage along the balance drum increases. This leads to an increase of the intermediate pressure prevailing in the annular chamber. Since the second axial face of the non-rotary part is exposed to the intermediate pressure prevailing in the annular chamber, the force acting on said second axial face increases and, hence, the non-rotary part of the additional balancing device moves towards the rotary part, whereby the width of the radial relief passage is reduced. 
     According to a preferred embodiment the non-rotary part comprises a third axial face, which is arranged between the first axial face and the second axial face with respect to the axial direction, wherein the third axial face is exposed to the same pressure as the first axial face during operation of the pump. This configuration renders possible that the additional balancing device with the rotary part and the non-rotary part reduces the leakage flow along the balance drum, but does not influence, at least not in a significant manner, the axial thrust compensation 
     Preferably, a ring-shaped sealing element is arranged at the radially outer surface of the non-rotary part, wherein the ring-shaped sealing element is arranged between the second axial face and the third axial face with respect to the axial direction. The ring-shaped sealing element seals the pressure difference between the pressure acting on the first and the third axial face on the one side, which is e.g. the discharge pressure, and the pressure acting on the second axial side, which is the intermediate pressure. 
     Furthermore, it is preferred that the ring-shaped sealing element has a sealing diameter, which equals the outer diameter of the balance drum. 
     Preferably, the non-rotary part is configured to be movable in the axial direction against the force of a spring element. 
     In this configuration with the spring element it is advantageous when the spring element is configured to push the non-rotary part in physical contact with the rotating part at standstill of the pump, so that the radial relief passage is closed. Thus, the spring element is designed as strong that it can press the non-rotary part of the additional balancing device in physical contact with the rotary part, at least as long as no hydraulic forces act upon the non-rotary part. Furthermore, the spring element is designed to be weak enough so that the hydraulic force acting on the non-rotary part during operation of the pump can move the non-rotary part in the axial direction against the force of the spring such that the radial relief passages is opened. 
     Furthermore, it is preferred that the contact faces with which the rotary part and the non-rotary part are in physical contact with each other are designed to withstand the friction during start-up or shutdown of the pump. 
     For this purpose the non-rotary part can comprise a stationary wear ring, which is configured such that the non-rotary part can physically contact the rotary part only with the stationary wear ring. 
     As an alternative or as a supplement the rotary part may comprise a rotary wear ring, which is configured such that the rotary part can physically contact the non-rotary part only with the rotary wear ring. 
     According to a preferred embodiment the additional balancing device and the spring element are configured to maintain a minimum width of the radial relief passage during operation of the pump. Thus, in particular the rotary part, the non-rotary part and the spring element are dimensioned and configured in such a manner that during operation of the pump a minimum width of the radial relief passage with respect to the axial direction is achieved, therewith considerably reducing the leakage flow of the fluid through the radial relief passage and the axial relief passage. In particular by the self-adjustment of the width of the radial relief passage it becomes possible to maintain the minimum width of the radial relief passage over the entire operating range of the pump. Thus, the efficiency of the pump can be considerably increased. 
     According to a preferred embodiment in particular for a single stage configuration of the pump or for a multistage configuration of the pump with an in-line arrangement of the impellers, the rotary part is arranged for being exposed to a pressure, which is at least essentially the same as the discharge pressure. This means with respect to the axial direction the additional balancing device is arranged adjacent to the last stage impeller in case of a multistage pump or to the only impeller in case of a single stage pump. 
     Furthermore, it is preferred that the suction pressure prevails at the low pressure location during operation of the pump. This is quite a simple design, because the balance line can just be connected to the inlet of the pump. 
     For many applications the pump can be configured as a multistage pump, wherein the hydraulic unit comprises at least a first stage impeller, and a last stage impeller, and optionally at least one intermediate stage impeller, with each impeller fixedly mounted on the pump shaft. 
     In particular when the pump is configured as a multistage pump with an in-line arrangement of the impellers it is preferred that the rotary part of the additional balancing device is arranged adjacent to the last stage impeller with respect to the axial direction. 
     In particular from the constructional point of view it is a preferred measure that the rotary part of the additional balancing device abuts against the balance drum. 
     Further advantageous measures and embodiments of the invention will become apparent from the dependent claims. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       The invention will be explained in more detail hereinafter with reference to the drawings. 
         FIG. 1  is a schematic cross-sectional view of an embodiment of a rotary pump according to the invention, and 
         FIG. 2  is a more detailed cross-sectional view illustrating a configuration of the balance drum and the additional balancing device. 
     
    
    
     DETAILED DESCRIPTION 
       FIG. 1  shows a schematic cross-sectional view of an embodiment of a rotary pump according to an embodiment of the invention, which is designated in its entity with reference numeral  1 . The pump  1  is designed as a centrifugal pump for conveying a fluid, for example a liquid such as water. 
     The rotary pump  1  comprises a pump housing  2  having an inlet  3  and an outlet  4  for the fluid to be conveyed. The inlet  3  is arranged on a suction side and receives the fluid having a suction pressure SP. The outlet  4  is arranged on a discharge side and discharges the fluid having a discharge pressure DP, wherein the discharge pressure DP is larger than the suction pressure SP. The pump  1  further comprises a hydraulic unit  5  for conveying the fluid from the inlet  3  the outlet  4  and for pressurizing the fluid from the suction pressure SP such that the fluid is discharged at the outlet  4  with the discharge pressure DP. In  FIG. 1  the flow of the fluid is indicated by the dashed arrows without reference numerals, 
     The hydraulic unit  5  comprises at least one impeller  51 ,  52 ,  53  for acting on the fluid. 
     The pump further comprises a pump shaft  6  for rotating each impeller  51 ,  52 ,  53  about an axial direction A. The axial direction A is defined by the axis of the pump shaft  6 . A direction perpendicular to the axial direction A is referred to as a radial direction. The pump shaft  6  extends from a drive end  61  to a non-drive end  62 . In this embodiment of the pump the drive end  61  of the pump shaft  6  is located outside of the pump housing  2  and can be connected to a drive unit (not shown) for driving the rotation of the pump shaft  6  about the axial direction A. The drive unit can comprise, for example, an electric motor. Each impeller  51 ,  52 ,  53  is mounted to the pump shaft  6  in a torque proof manner. 
     In the following description reference is made by way of example to an embodiment, which is suited for many applications, namely that the rotary pump  1  is configured as a multistage pump  1 , wherein the hydraulic unit  5  comprises a plurality of impellers  51 ,  52 ,  53 , namely at least a first stage impeller  51 , a last stage impeller  52 , and optionally at least one intermediate stage impeller  53 , with each impeller  51 ,  52 ,  53  fixedly mounted on the pump shaft  6 . The impellers  51 ,  52 ,  53  are arranged one after another on the pump shaft  6 . The reference numeral  51  designates the first stage impeller, which is arranged closest to the inlet  3  for receiving the fluid with the suction pressure SP. The reference numeral  52  designates the last stage impeller  52 , which is the impeller  52  closest to the outlet  4 . The last stage impeller  52  pressurizes the fluid such, that the fluid is discharged through the outlet  4  with the discharge pressure DP. The reference numeral  53  designates an intermediate stage impeller  53 . Each intermediate stage impeller  53  is arranged between the first stage impeller  51  and the last stage impeller  52  when viewed in the direction of increasing pressure. 
     The embodiment shown in  FIG. 1  has nine stages, i.e. the hydraulic unit  5  comprises the first stage impeller  51 , the last stage impeller  52  and seven intermediate stage impellers  53 . Of course, the number of nine stages has to be understood exemplary. The plurality of impellers  51 ,  52 ,  53  may be arranged in an in-line configuration as shown in  FIG. 1  or in a back-to-back configuration. In case of embodiments of the pump  1  as a single stage pump the hydraulic unit includes only one impeller constituting the first stage impeller  51  or the last stage impeller  52 , respectively. 
     The multistage rotary pump  1  shown in  FIG. 1  is designed as a horizontal pump, meaning that during operation the pump shaft  6  is extending horizontally, i.e. the axial direction A is perpendicular to the direction of gravity. The rotary pump  1  shown in  FIG. 1  is configured without an outer barrel casing, for example as a BB4 type pump. In other embodiments, the rotary pump  1  may be designed as a horizontal barrel casing multistage pump, i.e. as a double-casing pump. 
     It has to be understood that the invention is not restricted to the embodiment of the rotary pump  1 . In other embodiments, the rotary pump can be designed for example as a vertical pump, meaning that during operation the pump shaft  6  is extending in the vertical direction, which is the direction of gravity. 
     The rotary pump  1  comprises bearings on both sides of the hydraulic unit  5  (with respect to the axial direction A), i.e. the rotary pump  1  is designed as a between-bearing pump. A first radial bearing  81 , a second radial bearing  82  and an axial bearing  83  are provided for supporting the pump shaft  6 . The first radial bearing  81  is arranged adjacent to the drive end  61  of the pump shaft  6 . The second radial bearing  82  is arranged adjacent or at the non-drive end  62  of the pump shaft  6 . The axial bearing  83  is arranged between the hydraulic unit  5  and the first radial bearing  81  adjacent to the first radial bearing  81 . The bearings  81 ,  82 ,  83  are configured to support the pump shaft  6  both in the axial direction A and in a radial direction. The radial bearings  81  and  82  are supporting the pump shaft  6  with respect to the radial direction, and the axial bearing  83  is supporting the shaft  6  with respect to the axial direction A. The first radial bearing  81  and the axial bearing  83  are arranged such that the first radial bearing  81  is closer to the drive end  61  of the shaft  6 . Of course, it is also possible to exchange the position of the first radial bearing  81  and the axial bearing  83 , i.e. to arrange the first radial bearing  81  between the axial bearing  83  and the plurality of impellers  5 ,  51 , so that the axial bearing  83  is closer to the drive end  61  of the shaft  6 . 
     In other embodiments the axial bearing  83  may be arranged next to the second radial bearing  82 , i.e. next to the non-drive end  62  of the pump shaft  6 . In such embodiments the axial bearing  83  may be arranged between the hydraulic unit  5  and the second radial bearing  82  or between the second radial bearing  82  and the non-drive end  62  of the pump shaft  6 . 
     A radial bearing, such as the first or the second radial bearing  81  or  82  is also referred to as a “journal bearing” and an axial bearing, such as the axial bearing  83 , is also referred to as an “thrust bearing”. The first radial bearing  81  and the axial bearing  83  can be configured as separate bearings as shown in  FIG. 1 , but it is also possible that the first radial bearing  81  and the axial bearing  83  are configured as a single combined radial and axial bearing supporting the shaft both in radial and in axial direction. 
     Usually the bearings  81 ,  82 ,  83  are provided in separate bearing housings  84 ,  85 , which are fixedly connected to the pump housing  2 . The first radial bearing  81  and the axial bearing  83  are arranged in a first bearing housing  84  arranged adjacent to the drive end  61  of the pump shaft  6 . The second radial bearing  82  is provided in a second bearing housing  85  arranged adjacent to the non-drive end  62  of the pump shaft  6 . 
     All bearings  81 ,  82 ,  83  are preferably configured as antifriction bearings, such as ball bearings. Of course, it is also possible that some or all bearings  81 ,  82 ,  83  are configured as hydrodynamic bearings. 
     The rotary pump  1  further comprises two sealing devices, namely a first sealing device  86  for sealing the pump shaft  6  at the suction side adjacent to the first stage impeller  51  and the inlet  3 , and a second sealing device  87  for sealing the pump shaft  6  between the hydraulic unit  5  and the first bearing housing  84 . With respect to the axial direction A the first sealing device  86  is arranged between the hydraulic unit  5  an the second radial bearing  82 , and the second sealing device  87  is arranged between the hydraulic unit  5  and the axial pump bearing  83 . Both sealing devices  86 ,  87  seal the pump shaft  6  against a leakage of the fluid along the shaft  6  e.g. into the environment. Furthermore, by the sealing devices  86  and  87  the fluid can be prevented from entering the bearings  81 ,  82 ,  83 . Preferably, each sealing device  86 ,  87  comprises a mechanical seal. Mechanical seals are well-known in the art in many different embodiments and therefore require no detailed explanation. 
     The rotary pump  1  further comprises a balance drum  7  for at least partially balancing the axial thrust that is generated by the hydraulic unit  5  during operation of the rotary pump  1 . The balance drum  7  is fixedly connected to the pump shaft  6  in a torque proof manner and arranged between the hydraulic unit  5  and the drive end  61  of the pump shaft, more precisely, between the hydraulic unit  5  and the second sealing device  87 . The balance drum  7  defines a front side  71  and a back side  72 . The front side  71  is the side or the space facing the hydraulic unit  5 . The back side  72  is the side or the space facing the second sealing device  87 , i.e. the side or the space facing away from the hydraulic unit  5 . The balance drum  7  is surrounded by a stationary part  21 , so that an axial relief passage  73  is formed between the radially outer surface of the balance drum  7  and the stationary part  21 . The stationary part  21  is configured to be stationary with respect to the pump housing  2 . The axial relief passage  73  forms an annular gap between the outer surface of the balance drum  7  and the first stationary part  21  and extends in axial direction A from the front side  71  to the back side  72 . 
     The axial relief passage  73  is also referred to as “axial gap” or as “axial labyrinth”. The term “axial” designates that the relief passage  73  extends in the axial direction A, such that the fluid passing through said axial relief passage  73  flows in axial direction A, i.e. in a direction parallel to the pump shaft  6 . 
     Furthermore, a balance line  10  is provided connecting the back side  72  with a low pressure location. The low pressure location is a location, where during operation of the pump  1  a pressure prevails, which is smaller than the pressure at the front side  71 . Preferably the suction pressure SP prevails at the low pressure location. This is achieved in the embodiment of the pump  1  shown in  FIG. 1  by connecting the balance line  10  to the inlet  3 , so that the balance line  10  is in fluid communication with the inlet  10 . Thus, the balance line  10  constitutes a flow connection between the back side  72  and the pump inlet  3 . The balance line  10  may be arranged—as shown in  FIG. 1 —outside the pump housing  2 . In other embodiments the balance line  10  can be designed as internal line completely extending within the pump housing  2 . 
     According to embodiments of the invention, an additional balancing device  9  is arranged between the balance drum  7  and the hydraulic unit  5 . For a better understanding  FIG. 2  shows a more detailed cross-sectional view illustrating a configuration of the balance drum  7  and the additional balancing device  9 . 
     The additional balancing device  9  comprises a ring-shaped rotary part  91  fixedly connected to the pump shaft  6 , preferably in a torque proof manner, and a ring-shaped non-rotary part  92 , which is movable only in the axial direction A and which is secured against a rotational movement, in particular against a rotation about the axial direction A. The rotary part  91  has an axial face  911  facing the hydraulic unit  5 . In the embodiment of the pump shown in  FIG. 1  the axial face  911  of the rotary part  91  faces the last stage impeller  52 , and the axial face  911  is exposed to a pressure, which is at least essentially the same as the discharge pressure DP prevailing at the outlet  4  during operation of the pump  1 . Of course, due to smaller pressure losses caused by the fluid communication between the outlet  4  and the rotary part  91  the pressure prevailing at the axial face  911  of the rotary part  91  can be somewhat smaller than the discharge pressure DP. However, this small difference will be neglected hereinafter. 
     It has to be noted that the rotary part  91  is not necessarily arranged adjacent to the last stage impeller  52 . For example, in a multistage pump with a back-to-back arrangement of the impellers in the hydraulic unit, the rotary part of the additional balancing device can be arranged adjacent to an intermediate stage impeller of the hydraulic unit, namely this intermediate stage impeller which is arranged at the axial end of the hydraulic unit that is next to the balance drum. In such embodiments the pressure prevailing at the axial face of the rotary part is usually considerably smaller than the discharge pressure prevailing at the outlet of the pump, for example the pressure equals the suction pressure plus half the difference between the discharge pressure and the suction pressure. 
     As it can be seen both in  FIG. 1  and in  FIG. 2 , the rotary part  91  and the non-rotary part  92  are configured to overlap with respect to the radial direction, i.e. with respect to the direction perpendicular to the axial direction A. The non-rotary part  92  is configured to be movable in the axial direction A against the force of a spring element  95 , such that a radial relief passage  93  is open during operation of the pump  1 . The radial relief passage  93  extends in the radial direction between the rotary part  91  and the non-rotary part  92 . 
     It has to be noted that the spring  95  is not necessarily required. In other embodiments the axial position of the non-rotary part  92  or the movement of the non-rotary part  92  in axial direction A, respectively, may be determined by hydraulic forces only, or by the combination of hydraulic forces with for example gravitational forces, friction forces, magnetic forces or other forces. 
     The radial relief passage  93  is also referred to as “radial gap” or as “radial labyrinth”. The term “radial” designates that the relief passage  93  extends in the radial direction, such that the fluid passing through said radial relief passage  93  flows in radial direction, i.e. in a direction perpendicular to the pump shaft  6 . 
     In the embodiment shown in  FIG. 1  and  FIG. 2  the spring element  95  acting on the non-rotary part  92  rests on the stationary part  21  delimiting the axial relief passage  73 . The spring element  95  can comprise a helical spring or a disk spring or a spring collar or a spring washer or any other spring-elastic element, which is suited to exert a force on the non-rotary part  92 , which is directed in axial direction A towards the rotary part  91 . 
     When the non-rotary part  92  moves in the axial direction A, it moves relative to the stationary part  21 . A ring-shaped sealing element  99 , for example an O-ring, is provided between the stationary part  21  and the non-rotary part  92  for sealing therebetween. 
     At the front side  71  in front of the balance drum  7  an annular chamber  94  is provided between the balance drum  7  and the rotary part  91  of the additional balancing device  9 . On the one side, the balance drum  7  delimits the annular chamber  94  with respect to the axial direction A. On the other side the annular chamber  94  is delimited with respect to the axial direction A by the rotary part  91 . With respect to the radial direction the annular chamber  94  is delimited at the radially inner side by the rotary part  91  and at the radially outer side by the non-rotary member  92 . Thus, the annular chamber  94  extends between the rotary part  91  and the non-rotary part  92  with respect to the radial direction. With respect to the axial direction A the annular chamber  94  extends between the rotary part  91  and the balance drum  7 . 
     Preferably, as it is shown in  FIG. 2  the rotary part  91  of the additional balancing device  9  abuts against the balance drum  7 . The balance drum  7  has a recess at the front site  71 , which is configured to receive the end of the rotary part  91 . 
     The non-rotary part  92  comprises a first axial face  921  and a second axial face  922  delimiting the non-rotary part  92  with respect to the axial direction A, wherein the first axial face  921  is arranged to be exposed to the same pressure as the axial face of the rotary part  91  facing the hydraulic unit  5 , here namely the discharge pressure DP, and wherein the second axial face  922  is arranged to be exposed to the pressure prevailing in the annular chamber  94 . The non-rotary part  92  is interposed—with respect to the axial direction—between the rotary part  91  and the stationary part  21 . 
     The pressure prevailing in the annular chamber  94  during operation of the pump  1  is referred to as intermediate pressure IP. The intermediate pressure IP is smaller than the discharge pressure DP and larger than the suction pressure SP as will be explained hereinafter. 
     In the embodiment shown in  FIG. 2 , the non-rotary part  92  comprises a third axial face  923 , which is arranged between the first axial face  921  and the second axial face  922  with respect to the axial direction. The non-rotary part  92  is configured such, that the first axial face  921  and the third axial face  923  have the same outer diameter. The outer diameter of the second axial face  922  is smaller than the outer diameter of the first and the third axial face  921 ,  923 . 
     The outer diameter of the third axial face  923  is dimensioned such, that the third axial face  923  and the stationary part  21  overlap with respect to the radial direction, so that a ring-shaped chamber is formed between the stationary member  21  and the third axial face  923 . In said ring-shaped chamber the spring element  95  is arranged, wherein the spring element  95  rests both on the third axial face  923  and the stationary part  21 . 
     Furthermore, the outer diameter of the first and the third axial face  921  and  923  is dimensioned such that the third axial face  923  is exposed to the same pressure as the first axial face  921  during operation of the pump  1 , i.e. the fluid may flow from the high pressure side in front of the first axial face  921  into the ring-shaped chamber, where the spring element  95  is located. 
     The ring-shaped sealing element  99  is arranged at the radially outer surface of the non-rotary part  92  between the second axial face  922  and the third axial face  923  with respect to the axial direction A. Consequently, the ring-shaped sealing element  99  has a sealing diameter which equals—at least essentially—of the second axial face  922  of the non-rotary part  92 . In  FIG. 2  the reference numeral R denotes the radius R of the ring-shaped sealing element  99 . Thus, the sealing diameter of the ring-shaped sealing element  99  equals two times the radius R. 
     Preferably, the sealing diameter of the ring-shaped sealing element  99  equals the outer diameter of the balance drum  7  as it is shown in  FIG. 2 . The outer diameter of the balance drum  7  is given by the axial thrust that has to be generated by the balance drum  7  to at least partially compensate the hydraulic thrust generated by the rotating impellers  51 ,  52 ,  53  during operation of the pump  1 . Thus, in practice, for a specific application the outer diameter two times R of the balance drum  7  is determined depending on the required balancing forces that have to be generated by the balance drum  7 . When the outer diameter of the balance drum  7  has been determined, the additional balancing device  9  is configured such that the ring-shaped sealing element  99  at the non-rotary part  92  has a sealing diameter, which is at least essentially the same as the outer diameter of the balance drum  7 . 
     As already mentioned, the ring-shaped sealing element  99  may be for example a O-ring, which is arranged e.g. in a circumferential groove provided in the outer surface of the non-rotary part  92 . Of course, the ring-shaped sealing element  99  may also be configured as a metallic sealing element  99  or as a sealing element  99  made of a graphite compound material or a plastic, e.g. a thermoplastic polymer such as polyether ether ketone (PEEK) or polytetrafluorethylene (PTFE). The ring-shaped sealing element  99  may also be configured as a coating on the non-rotary part  92 . Furthermore, the ring-shaped sealing element may be configured as a labyrinth sealing. 
     The non-rotary part  92  comprises a stationary wear ring  923  arranged at the first axial face  921 . The stationary wear ring is configured such that the non-rotary part  92  can physically contact the rotary part  91  only with the stationary wear ring  923 . During operating conditions of the pump, when the rotary part  91  and the non-rotary part  92  are in physical contact with each other the stationary wear ring  923  ensures that the rotary part  91  and the non-rotary part  92  to withstand the friction. Such operating conditions occur for example during start-up or shutdown of the pump  1 . 
     As an alternative or as a supplement the rotary part  91  can comprise a rotary wear ring (not shown), which is configured such that the rotary part  91  can physically contact the non-rotary part  92  only with the rotary wear ring. 
     The stationary wear ring  923  and/or the rotary wear ring have a wear resistant surface, e.g. a coating or they are manufactured from a wear resistant material, which withstand the friction between the rotary part  91  and the non-rotary part  92 . An example for such a material is a thermoplastic polymer such as polyether ether ketone (PEEK). Another example is a graphite compound material. 
     As an alternative to providing the separate wear ring  923  it is also possible to configure the non-rotary part  92  in one piece, i.e. without a separate wear ring  923 . In such embodiments the entire non-rotary part  92  may consist e.g. of a metallic material or a graphite compound material or a plastic material such as PEEK or PTFE. In such embodiments also the ring-shaped sealing element  99  may be formed integrally, i.e. in one piece, with the non-rotary part  92 . 
     The spring element  95  is configured to push the non-rotary part  91  in physical contact with the rotary part  92  at standstill of the pump  1 , so that the radial relief passage  93  is closed. 
     Thus, at stillstand of the pump  1  the spring element  95  pushes the non-rotary element  92  against the rotary part  91 , so that the radial relief passage  93  is closed and the stationary wear ring  923  of the non-rotary part  82  is in physical contact with the rotary part  91 . 
     During start-up of the pump  1  the hydraulic force acting both on the non-rotary part  92 , more precisely on the first axial face  921  of the non-rotary part  92 , and on the rotary part  91 , more precisely on the axial face  911  of the rotary part  91 , increases until the pump  1  reaches its nominal speed. The hydraulic force pushes the non-rotary part  92  away from the rotary part  91  against the force of the spring element  95 , namely to the right according to the representation in  FIG. 1  and  FIG. 2 , so that the radial relief passage  93  opens between the rotary part  91  and the non-rotary part  92  of the additional balancing device  9 . The fluid flows through the radial relief passage  93 , into the annular chamber  94  at the front side  71  of the balance drum  7  and then through the axial relief passage  73  along the balance drum  7  to the back side  72  and into the balance line  10 . 
     When the start-up of the pump  1  is finished and the pump  1  has reached its nominal speed or the desired speed, the discharge pressure DP prevails at the outlet  4  as well as on the axial face  911  of the rotary part  91  and on the first axial face  921  of the non-rotary part  92 . At the inlet  3  as well as on the back side  72  the suction pressure SP prevails. Due to the opening of the radial relief passage  93 , the intermediate pressure IP prevails in the annular chamber  94 . The intermediate pressure IP is anywhere between the discharge pressure DP and the suction pressure SP. 
     By this combination of the radial relief passage  93  and the axial relief passage  73  the overall leakage flow through the additional balancing device  9  and along the balance drum  7  can be considerably reduced as compared to known balancing systems having a balance drum only. Although the overall leakage flow is considerably reduced the balancing action regarding the axial thrust is at least not considerably reduced as compared to known balancing systems. 
     The front side  71  is located in the annular chamber  94 , so that during operation of the pump  1  the axial surface of the balance drum  7  facing the front side  71  is exposed to the intermediate pressure IP prevailing in the annular chamber  94  Therefore, a considerably large pressure drop takes place over the balance drum  7 . At the back side  72  essentially the suction pressure SP prevails due to the balance line  10 . 
     Since the front side  71  is exposed essentially to the intermediate pressure IP, the pressure drop exists over the balance drum  7  so that the two axial faces delimiting the balance drum  7  with respect to the axial direction A are exposed to different pressures, namely one is exposed to the intermediate pressure IP prevailing in the annular chamber  94  and the other one is exposed to the suction pressure SP. This results in a force that is directed in axial direction A to the left side according to the representation in  FIG. 2 , therewith counteracting the axial thrust generated by the hydraulic unit  5  during operation of the pump  1 . 
     During operation of the pump  2  the radial relief passage  93  has a width W (see  FIG. 1 ). The width W, i.e. the extension of the radial relief passage  93  measured in the axial direction A is self-adjusting. As already explained, the first axial face  921  of the non-rotary part  92  is exposed to the same pressure as the axial face  911  of the rotary part  91  facing the hydraulic unit  5 , namely essentially the discharge pressure DP. The second axial face  922  of the non-rotary part is exposed to the intermediate pressure IP prevailing in the annular chamber  94 . Thus, the width W of the radial relief passage  93  depends on the difference between the discharge pressure DP times the first axial face  921  exposed to the discharge pressure DP and the intermediate pressure IP times the second axial face  922  exposed to the intermediate pressure IP. 
     If, during operation of the pump  1 , the discharge pressure DP acting of the first axial face  921  of the non-rotary part  92  increases, the width W of the radial relief passage  93  increases, meaning that the radial relief passage  93  becomes broader with respect to the axial direction A. Consequently the leakage flow through the radial relief passage  93  into the annular chamber  94  increases, whereby the resistance for the fluid flowing through the axial relief passage  73  along the balance drum  7  increases. This leads to an increase of the intermediate pressure IP prevailing in the annular chamber  45 . Since the second axial face  921  of the non-rotary part  92  is exposed to the intermediate pressure IP, the force acting on said second axial face  921  increases and, hence, the non-rotary part  92  of the additional balancing device  9  moves towards the rotary part  91 , whereby the width W of the radial relief passage  93  is reduced. In an analogous manner the with W of the radial relief passage  93  is self-adjusting in case the discharge pressure decreases. 
     The additional balancing device  9  and the spring element  95  are configured to maintain a minimum width W of the radial relief passage  93  during operation of the pump  1 . Thus, in particular the rotary part  91 , the non-rotary part  92  and the spring element  95  are dimensioned and configured in such a manner that during operation of the pump  1  a minimum width W of the radial relief passage  93  is achieved, therewith considerably reducing the leakage flow of the fluid through the radial relief passage  93  and the axial relief passage  73 . In particular, by the self-adjustment of the width W of the radial relief passage  93  it is possible to maintain the minimum width W of the radial relief passage  93  over the entire operating range of the pump  1 . Thus, the efficiency of the pump  1  may be considerably increased due to the reduction of the leakage flow.