Multi-stage turbomachine with lateral flow

Abstract

A multi-stage turbomachine (100) for processing a working fluid, provided with at least one lateral flow (30) of working fluid, and having a casing (14), a shaft (16) supported by bearings (15), an inlet (12) of a main flow (20) of the working fluid, an exhaust nozzle (17) of the main flow (20), a first impeller (7) and a second impeller (8) mounted cantilevered with respect to the bearings (15); the turbomachine having also a nozzle (1) for introducing the lateral flow (30), a spiral scroll (5) for imposing a degree of vorticity on the working fluid of the lateral flow (30) and contained within the casing (14), an annular channel (13) for transporting the lateral flow (30), a mixing section (10) in which the main flow (20) and the side lateral (30) join and mix.

Description

[0001] MULTI-STAGE TURBOMACHINE WITH LATERAL FLOW

[0002] DE S CR I P T I ON

[0003] Technical sector of the invention

[0004] The present invention relates to a multi-stage turbomachine and, in particular, to a gas inj ection system for a centri fugal compressor having the impellers cantilevered with respect to the shaft bearings . In the following, particular reference will be made to a multi-stage centri fugal compressor, but the invention is equally applicable to any multi-stage turbomachine for processing a working fluid .

[0005] Known technique

[0006] As is known, a compressor is an operating machine capable of increasing the pressure of a compressible fluid ( gas or vapors ) using mechanical energy . Among the various types of compressors used in industry there are the so-called centri fugal compressors , in which energy is trans ferred to the gas in the form of centri fugal acceleration due to rotation, generally imposed by a prime mover ( electric motor, steam turbine or gas turbine ) and related transmission mechanism, to a member called rotor . The rotor is made up of one or more bladed wheels , called impellers , rigidly connected to a shaft supported by bearings .

[0007] Centri fugal compressors can be equipped with a single impeller , in the so-called single-stage configuration, or with multiple impellers and, in this case , with reference to multi-stage compressors .

[0008] More precisely, each of the stages of a centri fugal compressor is usually made up of a duct for suction of the gas to be compressed, an impeller, capable of providing kinetic energy to the gas , and a duct for connecting an impeller to the next stage , the task of which is to convert the kinetic energy of the gas coming from the impeller into pressure energy .

[0009] In particular, these ducts downstream of the impellers are made up of a first section of exhaust duct from the impeller, called a di f fuser, a substantially U-shaped connection called " crossover" , and a second section of duct feeding into the next impeller, called a return channel .

[0010] Multi-stage compressors can also be characteri zed by the relative position of the bearings with respect to the dif ferent compression stages . In particular, the bearings can be positioned externally to the centrifugal stages in the so-called ' 'among the bearings' ' configuration or with centrifugal stages external to the bearing in the so-called ' 'cantilever' ' configuration.

[0011] Modern multi-stage centrifugal compressors used in the petrochemical industry can be designed with gas injection and / or extraction systems on intermediate stages, also called lateral flows. Some typical applications of these compressors are represented by machines used in heat pump and refrigerator cycles, which use high molecular weight gases that are injected at intermediate stages depending on process needs. Gas injection usually takes place by means of plenums or volutes obtained in the stator parts of the compressor, between two consecutive stages, in connection with an external flange. A compressor with such injection is also called a "lateral flow compressor".

[0012] Figure 1 shows an example of a compressor with lateral flow ("injection inlet " and " injection scroll ") and a ' 'between bearings' ' impeller configuration, while figure 2a shows an example of a compressor with lateral flow ("injection scroll") and ' 'cantilever' ' configuration of the impellers. Cantilever compressors can be coupled with a gear box, i . e . with a revolution multiplier and are called integrated multiplier compressors ( or IGC, an acronym from the English ' ' Integral Gear Compressor ' ' ) . In figure 2b, for example , a three- stage compressor is illustrated . Compressors with integrated multiplier are multi-shaft machines where the shafts that mount the compressor impellers have the pinion teeth made on the same shaft . The shafts and related pinions are arranged around a central gear connected to a primary shaft which is in turn connected to the driving machine that drives the compressor . In this type of compressor, it is possible to create a perfect combination between shaft speed and impeller si ze for each secondary shaft . All the impellers are cantilevered and can be equipped at the inlet with a plurality of adj ustable blades to guide the inlet flow . Inter-stage cooling of the gas flow can be carried out after each discharge from the impeller .

[0013] The combination of these features allows for high flow rates , exceptional energy ef ficiency even at partial load, all in a compact design .

[0014] Among the limitations imposed by the IGC architecture , the compactness requirement limits the possibility to use a large lateral flow with this configuration .

[0015] The IGC architecture , especially one with more than two impellers , requires to :

[0016] - limit the axial length of the stages to keep the cantilever mass relatively close to the bearings . An excessive distance between the impellers and the bearing makes the compressor unfeasible from the point of view of rotor dynamics with vibrations beyond the acceptable level ;

[0017] - limit the radial dimension of the compressor casing even in the presence of multiple compressors , i . e . limit the si ze of the central gear wheel . In fact , the diameter of the central gear wheel determines the distance between the pinions of the satellite gears and therefore between the axles of the various mounted compressors . To ensure that the external volutes of these compressors do not collide with each other, it would be advisable to use a large diameter central gear wheel but the si ze of this wheel is limited by the centri fugal forces that act on it depending on the diameter and the speed .

[0018] The current state of the art of IGC compressor technology makes it di f ficult to design multiple compressors connected to the same central gear in which at least one lateral flow is desired between two impellers mounted on the same pinion . This is due to the design limitation of the dimensions of the central gear wheel ( as stated above ) and the dimensions of the lateral inj ection system available according to the known technique . Figure 3 shows the state of the art for the geometry of the lateral flow inj ection system, both in terms of axial and radial si ze .

[0019] The so-called downstream inj ection geometry ( downstream means of the return channel ) is such as to increase the axial distance between the impeller and the bearing ( figure 3a ) .

[0020] The so-called "upstream inj ection geometry" limits the axial distance but presents the inj ection scroll on the external diameter of the compressor, thus increasing its radial si ze ( figure 3b ) .

[0021] There is , therefore , a need for a design solution for the lateral flow inj ection system which solves or at least mitigates the above-mentioned drawbacks .

[0022] Summary of the invention

[0023] An obj ect of the present invention is to improve the side flow inj ection system in a multistage turbo-machine provided with one or more lateral flows of the working fluid .

[0024] The invention is particularly, but not exclusively, dedicated to a centri fugal compressor in which the impellers are mounted cantilevered from the bearings supporting the rotating shaft of the compressor .

[0025] Even more particularly, the invention lends itsel f well to architectures with an arrangement of multiple centri fugal compressors connected to a single revolution multiplier according to the IGC ( Integrally gear compressor ) scheme , i . e . in the presence of a central gear wheel which drives a series of satellite pinions and therefore of compressors connected to them .

[0026] The invention aims to overcome the current limitations of compressor side flow inj ection technology to enable the design of an integral gear compressor with more than two stages housed around a common gear and large intermediate inj ection flows between the impellers .

[0027] In particular, the new lateral flow configuration provides , upstream of the first impeller, an inj ection screw and a suitable channel , bladed in some executions or without compartments in other executions , to connect the inj ection screw to the mixing section located upstream of the return channel (' 'upstream injection' ' ) . The channel that connects the injection screw, i.e. which distributes the fluid circumferentially, to the mixing section, is positioned externally, at a greater radial distance, compared to the connection curve between the diffuser and the return channel (' 'cross over' ' )

[0028] The position of the injection screw increases the overall axial length of the compressor but does not affect the overhang between the end impeller and the bearing as it is mounted 'upstream' of the first impeller, i.e. at the end of the shaft.

[0029] The injection screw is positioned as close as possible to the rotation axis of the compressor and therefore this component does not influence the radial overall dimensions of the compressor compared to an application without lateral flow injection. A minimal increase in radial overall dimensions is induced only by the presence of the channel that connects the screw to the mixing section. This channel, by virtue of the annular surface that develops over the entire external diameter at the ' 'cross over' ', has a reduced size, typically between 10 and 25% of the radial size compared to a screw positioned according to the known technique. Advantageously, upstream of a mixing section, a bladed array is provided, the blades of which are shaped in such a way as to control the tangential speed and the inclination of the lateral flow with minimal incidence losses at the entrance of the bladed distribution channel .

[0030] According to one aspect of the present invention, a multi-stage turbo-machine for the processing of a working fluid is therefore described, provided with at least one lateral flow and having the characteristics set out in the independent product claim attached to the present description .

[0031] Further preferred and / or particularly advantageous ways of implementing the aforementioned system are described according to the characteristics set out in the attached dependent claims .

[0032] Brief description of the drawings

[0033] The invention will now be described with reference to the attached drawings , which illustrate some non-limiting examples of implementation of the invention, in which : figure 1 schematically illustrates a compressor with lateral flow and configuration of the impellers between bearings , according to the known art , figure 2a schematically illustrates a compressor with lateral flow and cantilevered impeller configuration, according to the known art , figure 2b illustrates , in cross section, a compressor, equipped with an integrated revolution multiplier, according to the known art , figures 3a and 3b schematically illustrate two known configurations of the lateral flow inj ection system, according to the known art , figure 4 schematically illustrates a compressor with lateral flow and cantilevered impeller configuration, according to a preferred embodiment of the present invention, and

[0034] - figure 5 schematically illustrates a detail of the compressor in figure 4 .

[0035] Detailed description

[0036] With reference to figure 4 , a multi-stage turbo-machine 100 will be described which, according to a preferred embodiment of the invention, is a centri fugal compressor with cantilevered impellers and with lateral flow for which centrifugal compressor the same reference 100 will be used . Figure 4 partially schemati zes the compressor by omitting components that are not part of the present invention .

[0037] The compressor 100 , according to the example proposed, is a two-stage compressor provided with a lateral flow between the first and second stages . What is said below regarding the design of the components that allow the inj ection of the lateral flow inside the compressor must also be considered valid in the case of multi-stage compressors and, more generally, of any multi-stage turbo-machine , as the lateral flow inj ection can occur between any two consecutive stages of the compressor and / or turbomachine .

[0038] The compressor 100 is provided with a casing 14 and a rotor assembly which includes a first impeller 7 relating to the first compression stage and a second impeller 8 relating to the second compression stage . A shaft 16 of the compressor (partially illustrated) , which is rotatable around a rotation axis X, is supported by a pair of bearings 15 ( only one of which is illustrated in the figure , the other one being positioned at the right of the represented one ) , and the first impeller 7 and the second impeller 8 are both mounted cantilevered with respect to the pair of bearings 15 . The compressor is adapted to process a main flow 20 of a working fluid which enters the compressor by means of an inlet 12 according to an axial direction and, after being compressed in the first impeller 7 and subsequently in the second impeller 8 , it exits the compressor 100 in a radial direction by means of an exhaust duct 17 .

[0039] The new design involves an inj ection of lateral flow working fluid 30 , with the inj ection scroll positioned as close as possible to the rotation axis of the compressor so as not to influence the radial dimensions of the compressor nor the cantilevered impeller with respect to the bearings .

[0040] The lateral flow 30 of the compressor 100 , for example the flow coming from a separator or an evaporator at a pressure level intermediate between the suction and discharge of the compressor, feeds the compressor within which it will be mixed with the main flow coming from the first stage . The design must be such as to obtain low hydraulic losses , for example from 0 . 5 to 3% of the total pressure drop in the various ducts , while maintaining a minimum radial si ze , i . e . without impacting the design of the speed multiplier, and a minimum protrusion of the rotor ( i . e . the distance of the first impeller from the bearing) so to have a robust design from a rotor-dynamic point of view .

[0041] The new design is illustrated in figures 4 and 5 : the lateral flow 30 is directed into the compressor 100 and comes from a noz zle 1 provided with a flange 11 ; the noz zle 1 is preferably tangential to the casing 14 of the compressor 100 ( as seen in Fig . 5 ) to follow the shape of the subsequent components without causing excessive deviations of the gas flow . From the noz zle the lateral flow rate is distributed in a spiral scroll 5 which serves to uni formly feed the channel 13 and to impose a degree of vorticity on the gas .

[0042] A peculiarity of the new design compared to the known art is that the scroll 5 is located upstream of the di f fuser 6 of the first impeller 7 where the si ze of the scroll does not influence the design of the rotor, for example , the position of the impellers with respect to the bearing .

[0043] The space between the di f fuser 6 and the inlet 12 of the compressor 100 allows the scroll 5 to be si zed with a large cross section so as to have a speed lower than that which is typically obtained in known embodiments and preferably is not higher than 55 m / s , for example , between 30 and 55 m / s . The lower velocity has a positive impact on hydraulic losses which are proportional to the gas dens ity multiplied by its velocity squared . In applications such as heat pumps the molecular weight of the f luid is relatively high, ranging from 44 to 130 g -mol- 1 , so it is important to minimi ze the flow velocity as the density is directly proportional to the molecular weight of the gas .

[0044] The scroll 5 has a spiral shape with its section which progressively reduces circumferentially to maintain a constant gas velocity and therefore a uni form supply of the channel 13 . The shape of the scroll is such as to create a circumferential motion of the gas around the axis of rotation of the compressor 100 . This vorticity of the gas making up the lateral flow is necessary as the lateral flow follows a behavior similar to that of the gas coming from the first impeller 7 . It i s in fact well known the gas flow downstream a centri fugal impeller maintains a quantity of vortex which is trans formed into pressure by the return channel .

[0045] The scroll develops inside the casing 14 of the compressor 100 and does not influence the radial dimensions of the compressor, unlike what occurs in embodiments according to the known technique ( for example , that illustrated in Fig . 3b ) in which the scroll influences the radial dimensions of the compressor .

[0046] The scroll 5 supplies the annular channel 13 which is located in a radially external position but still adj acent to an elbow 4 , U-shaped and of small radial extension : this solution is such as to minimi ze the radial si ze of the entire compressor 100 , typically between 10 and 25% of the radial si ze of a screw positioned according to the known technique . The channel 13 is necessary to direct the lateral flow from the scroll 5 towards a mixing section 10 where it begins to mix with the main flow coming from the first impeller 7 ( the first stage impeller ) through the di f fuser 6 and the elbow 4 .

[0047] For the best fluid dynamic design of the mixing of the two flow rates it is necessary to match both the speed and the direction of the gas . A di f ference in speed or angle introduces an exchange of momentum associated, like any real trans formation, with an increase in pressure losses .

[0048] Considering a typical centri fugal compressor for heat pump and refrigeration service , the main flow in the mixing section 10 typically has a speed between 15 and 50 m / s and a direction with respect to the tangential direction that forms an angle between 15 ° and 30 ° . Ideally, the lateral flow at the mixing zone 10 should have the same velocity and slope angle values to have a uni form flow mixture with minimal losses .

[0049] The speed and angle of the lateral flow in the mixing zone 10 depend on the annular area of the channel 13 and on the tangential vortex that is created inside the scroll 5 . In particular, the speed in the channel 13 is inversely proportional to the cross-section of the channel while the flow angle , associated with the vortex, follows the well- known behavior of the ' ' free vortex ' ' with constant gas momentum . The gas momentum is the product of the tangential velocity times the distance of the section from the axis of rotation .

[0050] Since the new design obj ective minimi zes the velocity in the screw, and therefore the hydraulic losses , the vortex imposed on the lateral flow could be such as to create a di f ferent slope angle of the lateral flow than that of the main flow in the mixing section 10 . This implies , in other words , a lateral flow direction further away from the tangential direction . This angle di f ference , typically between 5° and 15°, cannot be compensated by changing the area of the channel 13, since the height of the channel 13 is set to control the adaptation of the velocity of the lateral flow relative to the main flow. In fact, it is known that the flow angle is given by the vector composition of the tangential component of the velocity and that normal to it, i.e. which characterizes the volumetric flow rate in the channel.

[0051] In cases where this condition occurs, a solution envisaged by the present invention is to install, upstream of the mixing section 10, a bladed array, which provides the required gas deflection. As illustrated in figure 4, it is possible to provide an array of axial blades 2 or an array of radial blades 3. Preferably, the number of blades is included in the range from 20 to 50 to provide the necessary 'solidity' (defined as the ratio between axial length of the bladed array and the relative circumferential pitch) of the bladed array.

[0052] The bladed array (both the axial blade array 2 and the radial blade array 3) could be integrated into the body of the channel 13 (for example, made by casting together with the diaphragm containing the blades) or assembled separately or it could be machined in a separate disk which is bolted to the body of the channel 13 . The shape of the blades is typically a circular arc but they can also be produced with more complex geometries with three- dimensional airfoil sections .

[0053] Both the main flow and the lateral flow after mixing in the mixing section 10 are directed into a bladed distribution channel 9 , which produces only a very small increase in the axial distance between the first impeller 7 of the first stage and the second impeller 8 of the second stage . Therefore , the almost negligible increase in the overhang of the first impeller 7 with respect to the bearing 15 is another advantage compared to the reali zations according to the known technique ( for example , that of figure 3a ) : in fact , a reduced overhang improves the rotor-dynamic behavior which allows to use a rotor with a shaft having a relatively small diameter . The reduction of the shaft diameter has a notable positive impact on performance since with a small shaft the flow entering the impeller has a lower relative speed and consequently lower associated losses . It is known that the relative velocity of the flow entering the impeller is the vector sum of the absolute velocity in the duct upstream of the impeller and the entrainment velocity proportional to the shaft diameter .

[0054] The distance traveled by the fluid between the mixing section 10 and the second impeller 8 is greater than the distance traveled by the fluid in some known embodiments ( for example , that of figure 3a ) . In this way it is possible to have a more uni form flow to the second impeller 8 with beneficial consequences for the aerodynamics of the impeller, for example by virtue of a flow with minimal variations in incidence at the entrance to its blades , and a lower intensity of disturbances of flow, for example wakes and turbulences , which can negatively influence the performance of the stator components of the second stage up to the exhaust noz zle 17 , for example compromising the ability to convert the pressure to the dynamic component of the flow .

[0055] In addition to the ways of implementing the invention, as described above , it should be understood that numerous further variations exist . It must also be understood that said methods of implementation are only exemplary and do not limit neither the obj ect of the invention, nor its applications , nor its possible configurations . On the contrary, although the above description makes it possible for the skilled man to implement the present invention at least according to one of its exemplary configurations , it must be understood that numerous variations of the described components are conceivable , without thereby departing from the obj ect of the invention, as defined in the attached claims .

Claims

CLAIMS1. Centrifugal compressor (100) for processing a working fluid, provided with at least one lateral flow (30) of working fluid, the compressor (100) comprising :- a casing ( 14 ) ,- a shaft (16) rotating around an axis of rotation (X) and supported by bearings (15) ,- an inlet (12) of a main flow (20) of the working fluid,- a discharge nozzle (17) of the main flow (20) ,- a rotor group comprising at least a first impeller (7) and at least a second impeller (8) , in which the first impeller (7) and the second impeller (8) are both cantilevered and on the same side with respect to the bearings (15) , the compressor being characterized in that, for the processing of the lateral flow (30) of working fluid, it further comprises the following components in combination:- a nozzle (1) for supplying the lateral flow (30) , a spiral scroll (5) to impose a degree of vorticity on the working fluid of the lateral flow (30) , contained within the casing (14) upstream of a diffuser (6) of the first impeller (7) ,- an annular channel (13) for carrying the lateral flow (30) located in a radially external position and adjacent to an elbow (4) , U-shaped and crossed by the main flow (20) .- a mixing section (10) in which the main flow (20) and the lateral flow (30) join and mix each other, said components being configured in such a way as not to influence either the distance between the first impeller (7) and the second impeller (8) with the corresponding bearings (15) , nor the radial dimensions of the compressor (100) .

2. Compressor (100) according to claim 1, wherein upstream of the mixing section (10) a bladed array (2, 3) is located, the array shaped for the purpose of imparting a predetermined inclination of the lateral flow (30) .

3. Compressor (100) according to claim 1, further comprising, downstream of the mixing section (10) a bladed distribution channel (9) , crossed by both the main flow (20) and the lateral flow (30) and configured so as to minimize the axial distance between the first impeller (7) and the second impeller (8) of the second stage, i.e. with a distance no greater than that obtainable in a compressor without lateral flow with a returnchannel which processes a volumetric flow rate equal to that resulting from the mixing of the main flow (20) and the lateral flow (30) in the mixing section (10) .

4. Compressor (100) according to claim 1, wherein the nozzle (1) is tangent to the casing (14) .

5. Compressor (100) according to claim 1, wherein the scroll (5) has a cross section such that the speed (v) of the lateral flow (30) is constant and does not exceed 55 m / s .

6. Compressor (100) according to claim 2, wherein the bladed array is an array of axial blades (2) .

7. Compressor (100) according to claim 2, wherein the bladed array is an array of radial blades ( 3 ) .

8. Compressor (100) according to claim 2, wherein the bladed array (2, 3) has a number of blades ranging from 20 to 50.

9. Compressor (100) according to claim 2, wherein the blades of the bladed array (2, 3) have a circular arc or three-dimensional airfoil shape.

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