FR2779771A1 - Turbo pump with multistage pump and turbine assemblies - Google Patents

Abstract

The multistage turbo pump is constructed within a casing and consists of a turbine and multistage pump. The rotor is mounted in the case such that it may be displaced axially by means of the elastic dampers, each of which consists of a socket (30), with an associated damper (31). The damper carries brackets (32, 33), on both the end faces, which are located to form a radial clearance (B) with the case. The socket and damper are also fitted with stops (35,36) designed to immobilize them separately.

Description

/ 1

TURBOPOMPE GROUP

  The present invention relates to installations for

  pumping and in particular relates to a turbopump group.

  It is more efficient to use the present invention for the production of liquid rocket boosters as well as in other branches of industry such as: aeronautical construction industry, oil and gas production and chemical industry or the we use the pumps

  and turbines operating under high loads.

  The life potential and the reliability of the operation of liquid rocket boosters depend to a large extent on a turbopump group. The life potential of a turbopump group depends on the lifetime of the bearings which depends on the axial and radial loads, acting on a rotor. The discharge of the bearings from the action of axial and radial forces is one of the most important solutions which are applied to ensure a high life potential and a high operating reliability of rocket boosters.

liquid.

  The axial discharge of the bearings of a rotor of a group

  turbopump is performed by axial discharge means.

  The discharge capacity of these means is chosen so as to ensure its highest value in order to effectively compensate for the large axial forces acting on the elements of the rotor during start-up and during the non-stationary operating regimes of a thruster. - The dynamic radial loads acting on a rotor depend on the quality of the balancing of a rotor during its manufacture, on the particular property of a rotor consisting in that during its period. of operation it does not cause the increase of the loads due to the initial imbalances and also of the capacity of damping of the supports of the rotoro

  We know a turbopump group (UK, A, 2083860) comprising

  both a multistage pump and a turbine formed by one, arter

and a rotor.

  The rotor has two bearings, each consisting of two bearings. The rotor is mounted in the casing by means of the elastic damping bearings of the bearings, which each include a bushing with a damper mounted thereon. The damper is produced in the form of a packet of metal blades. The impeller wheels of the pump and the turbine discs are mounted between the rotor supports on a shaft. The finned wheels are provided

  bearing surfaces which come into contact with the shaft.

  The rotor can move in the axial and radial directions.

  The axial stops of the bearings arranged on the ends of the shaft limit the movement of the rotor in the axial direction, The elastic damping supports limit

the radial movements of the rotor.

  The turbopump group is provided with a means for the automatic axial discharge of the rotor which consists of two chambers formed by a impeller of the last stage of the pump 3-

3- 2779771

  and the housing. Channels are made in the rotor to communicate one of said enclosures with an enclosure of

  low pressure. The other enclosure of said means is pooled

  cation with an enclosure located at the outlet of the impeller of the last stage of the pump. A seal and annular separation of

  the shaft is mounted between the pump and the turbine and is

  to separate the pump enclosure from the enclosure of the

  turbine. The sealing and separation packing is

  designed to prevent engine liquid from leaking from the pump enclosure to the enclosure between the casing and the turbine disc which applies to said seal and separation packing. The major drawback of the known turbopump group is that a very high axial force acts on the rotor of the group during the start-up and the operation of the rocket propellant.

  at low speed revolutions.

  The starting of a liquid rocket propeller is carried out at an excess of the power of the turbine compared to the power of the pump. As a result, the ratio of the different bressions observed at the turbine to the difference of

  pressures created by the pump believes which results in an imbalance

  significant liberation of the axial forces acting on the discs of

  the impeller and the impeller of the pump.

  During start-up and operation of a thruster-

  liquid fuel rocket at low speeds, a very large radial clearance is formed in the annular seal which connects the enclosure of the automatic axial discharge means of the ro-tor to the enclosure located - 4 - to the output of the fin wheel from the top stage by comparison

  with the main operating modes of a thruster-

  liquid rocket. This large radial clearance formed in the annular sealing gasket reduces the discharge capacity of the means.

automatic axial discharge.

  It follows that the efficiency of the axial discharge means

  rotor automatically descends during start-up and operation-

  operation of a liquid rocket propellant at low speeds of rotation when significant axial forces act

on the rotor.

  In a conventional construction of the turbopump unit, the inlet enclosure of a 2-impeller impeller of the last third stage of the pump serves as the low pressure enclosure. In this case, the difference in pressures created by a stage of the pump is used to adjust the axial force. way to

  automatic axial discharge of the rotor produced in accordance with this

  classic concept is characterized by a low discharge capacity. The increase in the discharge capacity of the automatic exial discharge means of the rotor is limited by the value of the surface area of the disk of the impeller of the pump. Low discharge capacity of the automatic axial discharge means of the rotor, the decrease in discharge efficiency during start-up and during operation of a liquid fuel rocket propellant at low revs and action high axial forces when starting the propellant on the rotor of the turbopump group causes the formation of high axial forces which act - 5- on the bearings and, therefore, their required service life cannot be guaranteed. To ensure proper operation

  bearings it is necessary either to use the limits

  start-up regimes and operating regimes

  ment of a propellant, or to use an auxiliary device.

  for example, a mechanical stop which would absorb during start-up

  rage of a thruster an axial force which has not been compensated by said automatic axial discharge means of the rotor. The elimination of these drawbacks leads to a complication in the construction of the means and to the increase in the cost price of a propellor. In addition, an essential drawback of the known turbopump group also lies in that the bearings are subjected to dynamic radial loads. considerable. These radial loads not only shorten the life of the bearings but also decrease the economic performance of the pump and the turbine. The higher the radial load, the greater the deflection of the rotor. Too high deflection of the rotor results in / wear of the packing element and contributes to the increased clearance in the packing and

  lowers the efficiency of the pump and the turbine.

  The significant dynamic radial loads are generated as a result of an imbalance of the rotor during the passage through the critical speeds of rotation. The rotors of hydrogen pumps are flexible, that is to say, their rotational speeds during operation are greater than the values of the critical speeds. The passage through critical rotational speeds is followed by high dynamic loads acting on the rotor. Under the action of these loads, the rotor shaft bends. At this time, the value of the deflection varies along the axis of the rotor. The maximum deflection value is located near the assembly surface of the impeller of the intermediate stage of the pump. Due to the fact that the impellers rest on the shaft at different places along the axis of said shaft and that the hubs of the impellers are rigid and have practically no deflection, the surfaces

  of the finned wheels move reciprocally.

  which results in an imbalance of the rotor.

  This phenomenon of unbalance of the rotor during passage through the critical rotational speeds takes place during tests of the rotor carried out several times during its balancing with rotation at high speeds. Therefore, it is impossible to obtain a good quality of the balancing of this rotor before the assembly of the turbopump group and during the operation of the propellant this rotor is characterized by high loads.

  dynamics acting on the bearings.

  of the rotor takes place during propulsion operation

  sister. In this case, at each start of the propellant, the values of the imbalance can be different because the values of the residual displacements of the assembly surfaces

  after stopping the thruster are different.

  A high imbalance of the rotor resulted from the assembly and an additional imbalance of the rotor formed during operation creates high dynamic radial loads acting on the bearings which it is impossible to ensure their required service life. In this case, the different - 7- copies of the trubopump group may differ significantly

  each other by the values of the dynamic loads acting

  health on the rotor and, therefore, by their life potentials. Dynamic radial loads acting on the rotor

  during the operation of the thruster depend on the characteristics

  of the damping of the rotor supports. Amortization

  ment assured in the supports significantly reduces the dynamic loads acting on the rotor. The damping provided in the supports is particularly effective in the case of flexible shafts during the operation of a rotor at

  speeds close to critical values.

  The elastic damping supports used in the

  known group are characterized by the following disadvantages.

  The value of the radial displacements of the rotor is established by means of the radial clearances in the support of the package of damper blades with a socket and a housing. The velocity of the games in a support is usually chosen taking into account the condition which imposes ensuring an admissible displacement of the rotor which eliminates the wear of the seals

  pump and turbine The value of these clearances is lower

  lower than the value of the games necessary to ensure the application

  cation of the damper in the radial direction to obtain the optimal characteristics of the damping of a support. Therefore, this reduced value imposed clearances in the support necessary for the conditions of normal operation of the seals of the pump and the turbine does not ensure the optimal characteristics of

  damping of a leaf damper.

  The bushing can also pivot during rotation of the rotor as a result of friction in the bearings together with the outer ring of the bearing and the damper blades can move reciprocally relative to each other along the circumference. To suppress the pivoting of the sleeve and the displacements of the blades of the shock absorber of the support, a stop is usually provided which cooperates with the sleeve and the blades of the shock absorber and with a cover. The stopper is subjected to the action of a circumferential force transmitted from the sleeve following the moment of friction in the bearing. The presence of a stopper on the cover results in a radial force acting on the bushing and, consequently, on the damper. This radial force harms the mobility of the support, leads to a tightening of the shock absorber in one direction and creates the conditions for the formation of different values of the characteristics of the elasticity and of the damping along the circumference. We therefore proposed to produce a turbopump group, the design of which would reduce the axial and radial forces

  dynamics acting on the bearings.

  _9 g _ 2779771 -9 - The problem is solved by the fact that the turbopump group comprising a casing and a rotor forming a multistage pump and a turbine with impellers and discs, mounted on a rotor shaft, a means of automatic axial discharge of the rotor, constituted by two cavities, formed by the impeller of the last stage of the pump and the casing, and by channels, produced in the rotor and placing one of said chambers in communication with a bass chamber pressure, the other enclosure of said means being connected to an enclosure located at the outlet of the impeller of the last stage of the pump, and a seal and annular separation, arranged between the pump and the turbine, the wheels with fins being provided with bearing surfaces and the rotor being mounted in the casing so that it can move axially on bearings by means of elastic damping supports which include ortent, each, a socket with a shock absorber, mounted on it, is characterized, according to the invention, by the fact that there is provided on the socket of each elastic damping support, on two sides of the end faces of the damper, of the flanges, which form jointly with the casing, a radial clearance, and by the fact that the bush and the damper are provided with retainers, intended to immobilize them separately. According to the proposed embodiment of the present invention, the radial clearance formed in the elastic damping supports limits the radial displacements of the rotor, said clearance formed by the flanges of the sleeve of the damping support and the casing, ensures the characteristics of elasticity and damping of the shock absorber thanks to the fact that it is possible to choose the required tightening of the shock absorber in the radial direction which is not

  limited by the permissible radial displacements of the rotor.

  -10 - During the operation of the turbopump group, the moment of friction in the bearings is greater than the moment of friction of the shock absorber. By using the retainers of the bushing and the shock absorber, the pivoting of the bushing has been eliminated and, consequently, the wear of the bushing and the damper caused by this pivoting. Thanks to this embodiment, the characteristics of elasticity and of

  the damping of the support during the operation of the propellant.

  0 The proposed design of the elastic damping support is characterized by high damping capacities and offers an effective means to increase the service life of the bearings and reduce the vibrations of the turbopump unit. Said support also reduces the radial loads acting on the rotor and, consequently, the deflection of the rotor at all operating speeds of the propellant including during passages at critical rotational speeds. Thanks to the reduction in the value of the deflection of the rotor, it is now possible to produce the seals for the pump and the turbine with reduced radial clearances, thereby improving the economic performance of the unit. These advantages are particularly important in the case of the use of flexible rotors. It is particularly important to monitor that the retainers of the socket are arranged regularly along the circumference of one of the

flanges.

  - 11- The moment of friction generated in the bearings is perceived by the stops of the excavation of the elastic damping support. Thanks to the fact that the retainers are arranged regularly along the circumference, the forces acting on them no longer create the radial load on the support. With this embodiment, it removes the radial displacements of the support caused by the moment of friction generated in the bearings and we managed to keep the characteristics of elasticity and damping

of support.

  The invention will be better understood and other aims, details and advantages of

  this will appear better in the light of the explanatory description which will

  follow with an - 12- of an embodiment given solely by way of nonlimiting example with reference to the attached nonlimiting drawings, in which: - Figure 1 shows an overall view of a turbopump group, in accordance to the invention, in longitudinal section; - Figure 2 shows a rotor of the turbopump group according to the invention, in longitudinal section;

  - Figure 3 shows an elastic damping support-

  pump tick in longitudinal section; - Figure 4 is a section along IV-IV in Figure 3; - Figure 5 shows an automatic axial discharge means of the rotor; - Figure 6 shows an auxiliary means for the automatic axial discharge of the rotoro The turbopump group comprises a casing 1 (Figure 1) and a rotor 2, which form a three-stage pump 3 and a two-stage turbine 4, a means 5 axial discharge

  automatic rotor 2 and auxiliary means 6 for de-

  automatic axial load. The rotor 2 rests on twin ball bearings 7, 8 and is mounted in a housing

  1 through the supports 9, 10 of elastic damping

  The rotor 2 is mounted so that it can move axially. The axial displacement of the rotor 2 is limited by axial stops 11, 12 of the casing 1o, -13 - The rotor 2 comprises a shaft 13. (Figure 2), on which are mounted the finned wheels 14, 15, 16 of the pump 3o The discs 17, 18 of the turbine 4 are associated with the shaft 13

  by a non-removable assembly. The fins 14, .16, of.

  first and last stages of the pump 3 each comprising two bearing surfaces 19, 20 and 21, 22 respectively. The

  cog wheel 14 is supported by its bearing surfaces 19,

  edged from the side of its inlet, co.tre a bearing surface 23 of the shaft 13. The impeller 16 is supported by its bearing surface 22 made on the side opposite its inlet against a bearing surface 24 of the shaft 13. The impeller 15 of the intermediate stage of the pump 3 rests on the side of the inlet against the bearing surface 20 of the fin wheel 14 and on the side opposite to the entry against the bearing surface 21 of the fin wheel 16. On a part between the bearing surfaces 19 and 22, the fin wheels 14, 15, 16 are mounted relative to the shaft 13 with a play annular "A" .o The value

  of play "A", that is, the difference between the interior radius

  the hubs of the finned wheels 14, 15, 16 and the radius

  uterior of the shaft 13 must be chosen greater than the

  2o maximum load on the shaft 13 and be within the limits of 0.05 mm to 3 mmo. The engine torque is transmitted from

  the shaft 13 auzx impellers 14, 15, 16 by l.'intermediate.

  diary of a splined assembly 25, 26, 27 * A bush 30 on the outer surface of which is mounted a damper 31 with blades is mounted in each elastic damping support 9, 10 of the pump 3 along the rings -14 - Outer 28, 29 (Figure 3) of the bearings 7, 8. Flanges 32, 33 are made on the side of the end faces on the sleeve of two sides of the damper 31. The outer surface of the flange 32 and the inner surface of the body 34 of each of the supports 9, 10 form a clearance "B" which limits, during operation, the value of the radial displacement of the sleeve 30 and, consequently, of the rotor 2. The value of the clearance "B" I is chosen account

  given the radial movements of the rotor 2 relative to the gar-

  sealing nitures of pump 3 and turbine 4 and is

  equal to 0.3 mm. A retainer 35 provided in the flange 32 on the side -

  of the shock absorber 32 eliminates the reciprocal displacement of the blades of the shock absorber 31 in the circular direction. Three retainers 36 of the socket 30 are regularly arranged on the flange 33 along its circumference and cooperate with grooves (not shown in the figure), produced in the body 34, The retainers 36 and the, cover 37 pocket the socket 30 from move axially The damper 31 with blades is mounted along the contact surfaces 38, 39 (FIG. 4) with a tightening with the sleeve 30 and the body 34. The value of the tightening along the contact surfaces 389 39 of the damper 31 must be chosen so as to ensure the required characteristics of elasticity and

amortization of supports 9, 10.

  The medium 6 for the automatic axial discharge of the rotor 2 comprises two chambers 40, 41 (FIG. 5) formed by the; fin wheel 16 and the housing 10 Channels 42, 43, 44 are formed in the hubs of the fin wheels 15, 16. The enclosure behind the disc of the fin wheel 15 is -15 _ separated from the brackets adjacent by seals 46, 47. The enclosure 48 located at the outlet of the impeller 16 is separated from the enclosure 40 by U.ne sealing liner 49. The enclosure 41 is separated from the adjacent enclosures by a seal 50 and by an annular projection 51 of the impeller 16, which is mounted relative to the casing

1 with an axial clearance "C".

  The channels 42 put the enclosure 41 in communication with the Game "A". Via the channels 44, the clearance "A" is connected to an inlet enclosure 52 of the impeller 15. The

  channels 43 put enclosure 45 in communication with Game "A".

  An annular projection 53 (Figure 6) is provided on the casing 1 and is placed relative to the end face of the disc 17 joined to it with a clearance "D" o A seal and separation. annular 54 installed between the pump 3 and the turbine 4 forms a joint with the projection an enclosure 55, of the auxiliary means 6. The radius of the projection

  53 is greater than the radius of the seal 54.

  A separation enclosure 56 is provided between the impeller 16 and the disc 17 of the turbine 4 and is connected by channels

  57 at the enclosure 48 located at the outlet of the impeller 16.

  The separation enclosure 56 is separated from the pump 3 by a seal 50 and from the turbine 4 by the seal

  sealing and annular separation 54.

  Before assembling the pump set, it is necessary to assemble the rotor 2 (figure 1) to make

- 16 -

  its balancing. To this end, the fin wheels 14 are mounted,

  , 16 and the bearings 7, 8 on the shaft 13 (Figure 2).

  Then, the rotor 2 is subjected to balancing within a required range of the speed of rotation during operation. After balancing, the rotor 2 is disassembled. Then, we proceed to the successive assembly of the casing 1 by mounting simultaneously,

in it the rotor 2.

  Before mounting, the axial supports 11, 12 have extra thicknesses for machiningo During mounting, the minimum and maximum values of clearance "C" are respectively ensured in

  the means 5 (FIG. 5) using the axial supports 11 and 12.

  We ensure the minimum and maximum values of the clearance "C" in displacement

  cutting the rotor 2 respectively to the axial supports 11 and 12. The difference between the maximum and minimum values of the

  jue "C" determines the value of the axial displacement of the rotor 2.

  To ensure, in the auxiliary means 6, the minimum value

  male of the clearance "D" (FIG. 6) between the annular projection 53 and the casing 1, the projection 53 is machined by moving the rotor 2 until the axial support 12. To ensure the possibility of machining during assembly, the annular projection 53 is produced before it is assembled with an extra thickness on its end face. The value of the clearance "D" is equal to the sum of the minimum value of this clearance and the value of the axial displacement of the rotor 20 To obtain, during assembly, a required value of the tightening according to the contact surfaces 38, 39 of the shock absorber 31 (FIG. 4), the external surface 30 is machined

- 17 -

  Before its assembly, the sleeve 30 has an additional thickness on the external surface necessary for the machining carried out during

its mounting.

  Before switching on the liquid fuel rocket propellant

  the chambers of pump 3 (figure 1) are filled with hydro-

  uncomfortable. Activated by the hydrogen pressure reigning in the enclosure of the pump 3, the rotor 2 moves until it is pressed.

  axial 11. Consequently, the minimum value of the clearance "C" (fi-

  gure 5) is ensured in medium 5 and the maximum value of

  clearance "D" settles in the means 6 (Figure 6).

  During the operation of the propellant, the hydrogen arrives at the inlet of the pump 3 and then passes through the impellers 14, 15, 16. A portion of hydrogen arrives from the enclosure 48 (FIG. 5) located the outlet of the impeller 16 to the enclosure 40 passing through the seal

  49 and then reaches through play "C" in enclosure '41.

  During the operation of the thruster, the value of the clearance "C" varies as a function of the axial position of the rotor 2. The pressure in the enclosure 40 varies as a function of the value of the clearance "C" and consequently, the value of l 'axial force acting on the wheel 2o0 fins 16 of the c8té of the enclosure 40 also varies. Through channels 42, clearance "A" and channels 44, the hydrogen is evacuated from enclosure 41 into the inlet enclosure 52 of the impeller 15. The hydrogen leaks arriving in 1 enclosure 45 through the sealing rings 46, 47 of the impeller are also evacuated in the inlet enclosure 52 through the channels 43, 44 and the clearance "A" o Line portion of hydrogen is evacuated from the enclosure

- 18 -

  48 (FIG. 6) situated at the outlet of the impeller 16 through the channels 57 towards the separation enclosure 56. The leaks of hyrogen from the separation enclosure 56 are evacuated through the seal 50 towards the enclosure 41 and through the seal and annular separation 54 towards the enclosure 55 of the auxiliary means 6. From the enclosure 55, the hydrogen reaches through the clearance "D" in the clearance "D "in the enclosure located downstream of disc 17 of the

turbine 4.

  When the propulsion is started and during operation

  i0 sor at main speeds, the axial forces which act on the discs 17, 18 of the turbine 4 in the direction of the pump 3 are balanced by the axial forces which act on the impeller 14, 15, 16 in the direction of the turbine 4. At the start-up and during a variation of the

  operating mode of the thruster, the axial forces acting on the discs 17, 18 and on the finned wheels 14, 15, 16

  unbalance and, as a result, the rotor 2 begins to move in the axial direction. The displacement of the rotor 2 causes the variation of the clearance "C" which results in the variation 2o of the pressure in the enclosure 40 and of the axial force, acting on the side of said enclosure 40 on the impeller 16. The movement of the rotor is interrupted after the increase in the force acting on the impeller 16 has balanced the axial force which caused the movement of the rotor 2. For 2 5 actuate the means 5, the difference of the pressures created by two stages of pump 3 due to the fact that the inlet enclosure

- 19 -

  52 of the impeller 15 of the pump serves as the low pressure enclosure. As a result, the middle 5 balances large

  axial forces acting on the rotor 2.

  The propellant is started at an excess of power from the turbine 4 relative to the power of the pump 3. As a result, the ratio of the difference in pressures at the turbine 4 to the difference in pressures created by the pump increases this which results in a significant imbalance of efforts

  axial acting on the discs 17, 18 and on the winged wheels

  lettes 14, 15, 16o Consequently, an axial force which the means cannot balance acts on the rotor 2. This is why,

  when the thruster starts, the rotor 2 continues to depress

  lace up to pump 3 after the discharge capacity of medium 5 has been fully used. Following the displacement of the

  rotor 2 in the direction of pump 3, the clearance "D" decreases.

  i5 When the rotor 2 is in the positions in which the clearance "D" is appreciably greater than the clearance in the belt

  sealing 54, the pressure in the enclosure 55 of the auxiliary means

  liaire 6 does not vary during movement of the rotor 2. After the clearance "D" has decreased to the value less than 0.3 mm, the pressure in the enclosure 55 begins to increase, which results in an increase in the force on the disc 17 acting in

  the opposite direction to the direction of movement of the rotor 2.

  The movement of the rotor 2 is interrupted after the increase in pressure in the enclosure 55 has balanced the axial force

  which has not been compensated by means 5.

  While the thruster is operating at operating speeds

  sation, the axial force to the turbine 4 decreases, the rotor 2 moves in the direction of the turbine 4 to a balanced position. Consequently, the clearance 'D "increases to a value greater than 0.3 mm and the pressure in the enclosure 55 no longer depends on the position of the rotor 2. After that? In the case of a variation of the operating mode of the propellant the imbalance resulting from this variation of the axial forces in the pump 3 and in the turbine 4 will be eliminated by the means 5o It follows that the auxiliary means 6 ensures an additional discharge of the axial forces acting at the time of starting For the operation of the device, leaks of liquids are used which arrive from the separation enclosure 56 and pass through the separation sealing gasket 54. Therefore, the auxiliary means 6 does not require any additional losses of the power needed to

its operation.

  Dynamic radial loads are generated as a result of an imbalance in the rotor 2. High dynamic radial loads act on the rotor 2 during operation at the rotational speeds of the rotor close to the critical rotational speeds 2o. The value of these loads depends on the damping properties of the elastic damping supports S9, 10. Under

  the action of radial loads, the shaft 13 moves jointly-

  ment with the bearings 7, 8 and with the sleeve 30 (figure 4) in the radial direction. Consequently, the damper 31 to 25 blades is tightened in the direction of movement by increasing the tightening along the contact surfaces 38, 39 in the -21 - direction of movement and decreasing the tightening in the opposite direction. effective amortization of charges

  dynamic is achieved in this case thanks to the forces of friction

between its blades.

  Under the action of a dynamic load generated by an imbalance, the rotor 2 rolls on the damper 31 in the circumferential direction while the retainer 35 suppresses the displacement

  blades in the circumferential direction of the shock absorber. 31.

  Dynamic load and moment of friction in bearings

  with balls 7, 8 create in the socket 30, a torque which contributes

  drank at its rotation. The retainers 36 prevent the socket 30

  to turn. Owing to the fact that the retainers 36 are arranged regularly

  lier on the flange 33, the forces acting on them are identical which eliminates the displacement of the sleeve under the action of forces acting on the retainers 36.t allows to maintain high damping properties of the

shock absorber 31.

  During thruster operation, the radial forces

  can act on the supports at 9, 10 for a short time and.

  cause a radial displacement of the rotor 2 which is greater than 2o0 clearances in the seals of the pump 3 and of the turbine 4o In this case, the risk of wear of the seals and, consequently, the decrease characteristics of the pump 3 and of the turbine 4 are removed because the radial displacements of the rotor 2 are limited by the play

  "B" between the flange 32 of the sleeve 30 and the body 34.

  The low rigidity of the damper blades 31 ensures

  during assembly and during operation, the compensation

- 22 -

  tion of the mutual displacements of the supports 9, 10. under the action of small forces acting on the bearings 7, 8o Thanks to the reduction in the displacements of the shaft 13 in the supports 9, 10 which have been obtained using of a high damping capacity of the damper 31 and by the play which limits the radial displacements of the supports 9, 10, the values of the plays in the seals of the pump 3 and of the turbine 4 are reduced this which ensures high economic performance for pump 3 and turbine 4o -23-

Claims (2)

  1.- Turbopump unit, comprising a casing (1) and a rotor (2) which form a multi-stage pump (3) and a turbine (4) with impellers (14, 15, 16) and discs (17, 18), mounted on a shaft (13) of the rotor (2), a means (5) of automatic axial discharge of the rotor (2) constituted by two chambers (40, 41), formed by the impeller (16) of the last stage of the pump (3) and by the casing (1), and by channels (42, 44) produced in the rotor (2) and putting one of said speakers (41) into communication with a 0io bass speaker pressure, the other enclosure (40) of said means (5) being connected to an enclosure (48) situated at the outlet of the impeller (16) of the last stage of the pump (3), and a seal and of annular separation (54), disposed between the pump (3) and the turbine (4), the impellers (14, 15, 16) being provided with bearing surfaces and the rotor (2) being mounted in the casing (1) so that it can move r axially on bearings (7, 8) by means of elastic damping supports (9,) which each comprise a bushing (30) with a damper (31) mounted thereon, characterized by fact that there is provided on the sleeve (30) of each elastic damping support (9, 10), on two sides of the end faces of the damper (31), flanges (32, 33), which form together with the casing (1), a radial clearance (B) and by the fact that the bush (30) and the shock absorber (31) are provided with retainers (35, 36) intended for them immobilize separately.   25.2.- Turbopump unit according to claim 1, characterized in that the retainers (36) of the sleeve (30) are regularly arranged according to the   circumference of one of the flanges (33).