Multi-stage turbine with lateral flow

Abstract

A multi-stage turbine (100) for treating a working fluid equipped with at least one lateral flow (30) of the working fluid and having a housing (14), a shaft (16) supported by a bearing (15), an inlet (12) for the main flow (20) of the working fluid, an exhaust nozzle (17) for the main flow (20), and a first impeller (7) and a second impeller (8) mounted cantilevered relative to the bearing (15); the turbomachine also has a nozzle (1) for introducing the lateral flow (30), a spiral volute (5) for imparting a degree of vorticity to the working fluid of the lateral flow (30) and contained in the housing (14), an annular channel (13) for conveying the lateral flow (30), and a mixing section (10) in which the main flow (20) and the lateral flow (30) merge and mix.

Description

Technical Field

[0001] The present invention relates to a multi-stage turbine, and more particularly to a gas injection system for a centrifugal compressor having an impeller cantilevered relative to a shaft bearing. While specific reference will be made below to a multi-stage centrifugal compressor, the present invention is equally applicable to any multi-stage turbine for processing a working fluid. Background Art

[0002] As we all know, a compressor is an operating machine that uses mechanical energy to increase the pressure of a compressible fluid (gas or vapor). Among the various types of compressors used in industry are so-called centrifugal compressors, in which energy is transferred to the gas in the form of centrifugal acceleration, which is usually caused by the rotation applied to a component called a rotor by a prime mover (an electric motor, steam turbine, or gas turbine) and an associated transmission mechanism. The rotor consists of one or more bladed wheels (called impellers) rigidly connected to a shaft supported by bearings.

[0003] Centrifugal compressors can be equipped with a single impeller, a so-called single-stage configuration, or with multiple impellers, in which case they are called multi-stage compressors.

[0004] More specifically, each stage of a centrifugal compressor generally consists of a suction pipe for the gas to be compressed, an impeller capable of providing kinetic energy to the gas, and a connecting pipe connecting the impeller to the next stage, whose function is to convert the kinetic energy of the gas from the impeller into pressure energy.

[0005] Specifically, these ducts downstream of the impeller consist of a first portion of the impeller's discharge duct (called the diffuser), a roughly U-shaped connection (called the "cross over"), and a second portion of the duct leading to the next impeller (called the return channel).

[0006] Multi-stage compressors can also be characterized by the relative position of the bearings with respect to the different compression stages. Specifically, the bearings can be located outside the centrifugal stages in a so-called "between-bearings" configuration, or the centrifugal stages can be located outside the bearings in a so-called "cantilever" configuration.

[0007] Modern multi-stage centrifugal compressors used in the petrochemical industry can be designed with gas injection and / or extraction systems located at the intermediate stages, also known as sideflow. Typical applications for these compressors are machines used in heat pumps and refrigeration cycles, which utilize high-molecular-weight gases that are injected at the intermediate stages according to process requirements. Gas injection is typically achieved via a plenum or volute, located in the compressor stator between two consecutive stages and connected to an external flange. Compressors with this injection feature are also known as "sideflow compressors."

[0008] Figure 1 shows an example of a compressor with lateral flow ("jet inlet" and "jet volute") and a "between bearings" impeller configuration, while Figure 2a shows an example of a compressor with lateral flow ("jet volute") and a "cantilever" configuration of the impeller.

[0009] Overhung compressors can be coupled to a gearbox (i.e., a speed multiplier), known as an integrated gear compressor (or IGC, short for "Integral Gear Compressor" in English). For example, Figure 2b shows a three-stage compressor. A compressor with an integrated multiplier is a multi-shaft machine in which the shafts on which the compressor impellers are mounted are formed with pinion teeth on the same shaft. These shafts and associated pinions are arranged around a central gear, which is connected to the main shaft, which in turn is connected to the drive machine that drives the compressor. In this type of compressor, a perfect combination can be achieved between the shaft speed of each secondary shaft and the impeller size. All impellers are cantilevered and can be equipped with multiple adjustable blades at the inlet to guide the inlet flow. Interstage cooling of the airflow can be performed after each discharge from the impeller.

[0010] The combination of these features allows for high flow rates, extremely high energy efficiency even at part load, and an overall compact design.

[0011] Among the limitations of the IGC structure, compactness requirements limit the possibility of using large lateral flows in this configuration.

[0012] IGC structures, especially those with more than two impellers, require:

[0013] -Limiting the axial length of these stages to keep the cantilevered masses relatively close to the bearings. Too large a distance between the impeller and the bearings would make the compressor rotor-dynamically unviable and vibrations would exceed acceptable levels;

[0014] Even with multiple compressors, the radial dimensions of the compressor housing, and therefore the size of the sun gear, must be limited. The diameter of the sun gear effectively determines the distance between the pinions of the satellite gears, and thus the distance between the shafts of the individual installed compressors. To ensure that the outer volutes of these compressors do not collide with each other, a large-diameter sun gear is preferred. However, the size of this gear is limited by the centrifugal forces acting on it, which depend on its diameter and rotational speed.

[0015] Prior art IGC compressor technology makes it difficult to design multiple compressors connected to the same sun gear, where at least one sidestream flow is desired between two impellers mounted on the same pinion. This is due to design limitations on sun gear size (as discussed above) and the size limitations of sidestream injection systems available according to prior art. FIG3 illustrates the prior art geometry of a sidestream injection system (in terms of axial and radial dimensions).

[0016] The so-called downstream injection geometry (downstream refers to downstream of the return flow channel) increases the axial distance between the impeller and the bearing (Figure 3a).

[0017] The so-called “upstream injection geometry” limits the axial distance, but the injection volute is located on the outer diameter of the compressor, thereby increasing its radial size (Figure 3b).

[0018] Therefore, there is a need for a design of a side stream injection system that solves or at least alleviates the above-mentioned disadvantages. Summary of the Invention

[0019] An object of the present invention is to improve a side stream injection system in a multi-stage turbine provided with one or more side streams of a working fluid.

[0020] The invention has particular, but not limited, use with centrifugal compressors in which the impeller is cantilevered from bearings supporting the rotating shaft of the compressor.

[0021] More specifically, the invention is well suited for structures having an arrangement of multiple centrifugal compressors connected to a single speed multiplier according to the IGC (Integrally Geared Compressor) scheme, i.e. there is a central gear which drives a series of satellite pinions and therefore the compressors connected to them.

[0022] The present invention aims to overcome the existing limitations of compressor sidestream injection technology to enable the design of an integrally geared compressor having more than two stages housed around a common gear and with a large intermediate injection flow between the impellers.

[0023] Specifically, the new lateral flow configuration includes a jet spiral upstream of the first impeller and a suitable passage (with blades in some embodiments or without a partition in other embodiments) connecting the jet spiral to a mixing section upstream of the return channel ("upstream injection"). The passage connecting the jet spiral (i.e., which distributes the fluid circumferentially) to the mixing section is positioned outboard and at a greater radial distance than the connecting curve ("transition section") between the diffuser and the return channel.

[0024] The location of the injection spiral increases the overall axial length of the compressor, but does not affect the overhang between the terminal impeller and the bearing, since it is mounted "upstream" of the first impeller, ie at the end of the shaft.

[0025] The injection spiral is positioned as close as possible to the compressor's axis of rotation, so it does not impact the compressor's overall radial dimensions compared to applications without sidestream injection. This minimal increase in overall radial dimensions is solely due to the presence of a channel connecting the spiral to the mixing section. This channel, thanks to an annular surface extending over the entire outer diameter at the "transition section," is reduced in radial dimension, typically by 10% to 25%, compared to spirals positioned according to the prior art.

[0026] Advantageously, a blade array is provided upstream of the mixing section, the blades of the blade array being shaped so as to control the tangential velocity and inclination of the lateral flow with minimal incidence losses at the inlet of the bladed distribution channel.

[0027] According to one aspect of the invention, there is therefore described a multi-stage turbine for processing a working fluid, the turbine having at least one side flow and having the features set out in the independent product claim appended to this description.

[0028] Further preferred and / or particularly advantageous embodiments of the above-described system are described in accordance with the features set out in the accompanying dependent claims. BRIEF DESCRIPTION OF THE DRAWINGS

[0029] The present invention will now be described with reference to the accompanying drawings, which show some non-limiting implementation examples of the invention, in which:

[0030] FIG1 schematically shows a compressor according to the prior art having a side flow and an impeller located between bearings;

[0031] FIG2 a schematically shows a compressor having a side flow and a cantilevered impeller configuration according to the prior art;

[0032] FIG2 b shows in cross section a compressor according to the prior art equipped with an integrated speed multiplier;

[0033] 3a and 3b schematically illustrate two known configurations of a lateral flow injection system according to the prior art;

[0034] Figure 4 schematically illustrates a compressor having a side flow and cantilever impeller configuration according to a preferred embodiment of the present invention; and

[0035] Figure 5 Schematically shows Figure 4 Details of the compressor. DETAILED DESCRIPTION

[0036] refer to Figure 4, a multi-stage turbine 100 will be described. According to a preferred embodiment of the present invention, the turbine is a centrifugal compressor with a cantilevered impeller and a side flow, and for this purpose the same reference numeral 100 will be used to refer to the centrifugal compressor. Figure 4 The compressor is shown partially schematically by omitting components that are not part of the invention.

[0037] According to the example presented, the compressor 100 is a two-stage compressor provided with a side flow between the first and second stages. The following description of the component design that allows the injection of a side flow into the interior of the compressor is equally applicable to multi-stage compressors and, more generally, to any multi-stage turbine, since the side flow injection can occur between any two consecutive stages of the compressor and / or turbine.

[0038] Compressor 100 includes a housing 14 and a rotor assembly including a first impeller 7 associated with a first compression stage and a second impeller 8 associated with a second compression stage. A compressor shaft 16 (partially shown) is rotatable about an axis of rotation X and is supported by a pair of bearings 15 (only one of which is shown; the other is located to the right of the illustrated bearing). Both first impeller 7 and second impeller 8 are cantilevered relative to the pair of bearings 15. The compressor is adapted to process a main flow 20 of a working fluid, which enters the compressor axially through an inlet 12 and, after being compressed in the first impeller 7 and subsequently in the second impeller 8, exits compressor 100 radially through a discharge conduit 17.

[0039] The new design involves the injection of a side-flow working fluid 30 with the injection volute positioned as close as possible to the axis of rotation of the compressor so as not to affect the radial dimensions of the compressor nor the cantilevered impeller relative to the bearings.

[0040] The side stream 30 of the compressor 100 (e.g., a stream from a separator or evaporator, with a pressure level between the suction and discharge pressures of the compressor) is fed into the compressor where it is mixed with the main stream from the first stage. The design must achieve low hydraulic losses (e.g., 0.5% to 3% of the total pressure drop in each pipe) while maintaining a minimum radial dimension (i.e., not affecting the design of the speed multiplier) and a minimum protrusion of the rotor (i.e., the distance of the first impeller from the bearing) in order to achieve a robust design from a rotor dynamics perspective.

[0041] Figure 4 and Figure 5 The new design is shown: the lateral flow 30 is directed into the compressor 100 and comes from a nozzle 1 provided with a flange 11; the nozzle 1 is preferably tangential to the casing 14 of the compressor 100 (e.g. Figure 5The lateral flow is distributed from the nozzle into the spiral volute 5, which is used to evenly feed the channel 13 and impose a certain degree of vorticity on the gas.

[0042] Compared with the prior art, a feature of the new design is that the volute 5 is located upstream of the diffuser 6 of the first impeller 7. In this position, the size of the volute does not affect the design of the rotor, such as the position of the impeller relative to the bearing.

[0043] The space between the diffuser 6 and the inlet 12 of the compressor 100 allows the volute 5 to be dimensioned with a large cross-section, thereby enabling a velocity lower than that typically achieved in prior art embodiments, and preferably no higher than 55 m / s, for example, between 30 and 55 m / s. Lower velocities have a positive impact on hydraulic losses, which are proportional to the density of the gas multiplied by the square of its velocity. In applications such as heat pumps, the molecular weight of the fluid is relatively high, ranging from 44 to 130 g·mol-1. Since the density of the gas is directly proportional to the molecular weight, minimizing the flow velocity is crucial.

[0044] The volute 5 has a spiral shape, its cross-section gradually decreasing circumferentially, in order to maintain a constant gas velocity and thus provide a uniform supply to the channel 13. The shape of the volute is designed to induce a circular motion of the gas around the axis of rotation of the compressor 100. This vortex of the gas constituting the lateral flow is necessary because the lateral flow follows characteristics similar to those of the gas from the first impeller 7. In fact, it is well known that the gas flow downstream of a centrifugal impeller maintains a certain amount of vortex, which is converted into pressure by the return flow channel.

[0045] The volute extends inside the casing 14 of the compressor 100 and does not affect the radial dimensions of the compressor, unlike embodiments according to the prior art (such as the embodiment shown in FIG. 3 b ) where the volute affects the radial dimensions of the compressor.

[0046] The volute 5 feeds an annular channel 13 located in a radially outer position, but still adjacent to the elbow 4, which is U-shaped and has a small radial extension: this solution minimizes the radial dimensions of the entire compressor 100, typically 10% to 25% of the radial dimensions of the spiral arranged according to the prior art. The channel 13 is required to guide the lateral flow from the volute 5 towards the mixing section 10, where it begins to mix with the main flow coming from the first impeller 7 (the impeller of the first stage) and passing through the diffuser 6 and the elbow 4.

[0047] To achieve the best fluid dynamics for mixing the two flows, both the velocity and direction of the gases need to be matched. As with any practical conversion, differences in velocity or angle result in an exchange of momentum and an accompanying increase in pressure loss.

[0048] Considering typical centrifugal compressors used in heat pumps and refrigeration equipment, the velocity of the main flow in the mixing section 10 is generally between 15 and 50 m / s, and the direction of the main flow forms an angle of between 15° and 30° relative to the tangential direction. Ideally, the lateral flows at the mixing zone 10 should have the same velocity and inclination angle values ​​to achieve uniform flow mixing with minimal losses.

[0049] The rate and angle of the lateral flow in mixing zone 10 depend on the annular area of ​​passage 13 and the tangential vortex generated inside volute 5. Specifically, the velocity in passage 13 is inversely proportional to the cross-section of the passage, while the flow angle associated with the vortex follows the well-known "free vortex" characteristic of constant gas momentum. Gas momentum is the product of the tangential velocity and the distance of the cross-section from the axis of rotation.

[0050] Since the goal of the new design is to minimize the velocity in the spiral section (and therefore minimize the hydraulic losses), the vortex imposed on the lateral flow may cause the inclination angle of the lateral flow in the mixing section 10 to be different from the inclination angle of the main flow. In other words, this means that the lateral flow direction is further away from the tangential direction. This angular difference (typically 5° to 15°) cannot be compensated by changing the area of ​​the channel 13, because 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 synthesis of the tangential component of the velocity and the component perpendicular to the tangential component (i.e., it characterizes the volume flow in the channel).

[0051] When this happens, the solution proposed by the present invention is to install an array of blades upstream of the mixing section 10, which provides the required gas deflection. Figure 4 As shown, there may be provided an axial array of blades 2 or a radial array of blades 3. Preferably, the number of blades is comprised within the range of 20 to 50 to provide the desired "solidity" of the blade array (defined as the ratio between the axial length of the blade array and the relative circumferential pitch).

[0052] The blade arrays (both the axial blade array 2 and the radial blade array 3) may be integrated into the body of the channel 13 (e.g., by casting them together with a diaphragm containing the blades), or assembled separately, or they may be machined into a separate disc that is bolted to the body of the channel 13. The shape of the blades is typically an arc of a circle, but they can also be made with more complex geometries having a three-dimensional airfoil cross-section.

[0053] Both the main flow and the lateral flow, after mixing in the mixing section 10, are directed into the bladed distribution channel 9, which only slightly increases the axial distance between the first impeller 7 of the first stage and the second impeller 8 of the second stage. Consequently, the almost negligible increase in the overhang of the first impeller 7 relative to the bearing 15 compared to prior art embodiments (e.g., the embodiment of FIG. 3 a ) is a further advantage: the reduced overhang improves the rotor dynamics, allowing the use of rotors with relatively small shaft diameters. A reduction in shaft diameter has a significant positive impact on performance, since with smaller shaft diameters, the flow entering the impeller has a relatively lower flow rate and, therefore, lower associated losses. As is well known, 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, which is 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 that traveled by the fluid in some prior art embodiments (e.g., the embodiment of FIG3 a ). In this way, the flow entering the second impeller 8 is more uniform, which has a beneficial effect on the aerodynamic performance of the impeller, for example, because the inclination angle of the flow at its blade inlet varies minimally and the intensity of flow disturbances (e.g., wakes and turbulence) is lower, which could negatively affect the performance of the stator components of the second stage up to the discharge nozzle 17 (e.g., impair the ability to convert pressure into the dynamic component of the flow).

[0055] In addition to the embodiments of the present invention described above, it should be understood that many other variations exist. It should also be understood that the methods of the above-described embodiments are merely exemplary and do not limit the purpose of the present invention, its applications, or its possible configurations. On the contrary, although the above description enables those skilled in the art to implement the present invention according to at least one exemplary configuration of the present invention, it should be understood that many variations can be made to the described components without departing from the purpose of the present invention as defined in the appended claims.

Claims

1. A centrifugal compressor (100) for processing a working fluid, said centrifugal compressor being provided with at least one lateral flow (30) of the working fluid, said centrifugal compressor (100) comprising: - housing (14); - a shaft (16) rotating around an axis of rotation (X) and supported by bearings (15); - an inlet (12) for the main flow (20) of the working fluid; - a discharge nozzle (17) of the main flow (20); - a rotor assembly comprising at least one first impeller (7) and at least one second impeller (8), wherein the first impeller (7) and the second impeller (8) are cantilevered and located on the same side relative to the bearing (15); The centrifugal compressor is characterized in that, in order to process the lateral flow (30) of the working fluid, the centrifugal compressor further comprises the following components in combination: - a nozzle (1) for supplying said lateral flow (30); - a spiral volute (5) for imparting a certain degree of swirl to the working fluid of the lateral flow (30), the spiral volute being contained in the housing (14) and being located upstream of the diffuser (6) of the first impeller (7); - an annular channel (13) for conveying the lateral flow (30), said annular channel being located in a radially outer position and adjacent to the U-shaped bend (4) passed by the main flow (20); a mixing section (10) in which the main flow (20) and the lateral flow (30) merge and mix with each other; These components are configured so as not to affect the distances between the first impeller (7) and the second impeller (8) and the corresponding bearings (15), nor to affect the radial dimensions of the compressor (100).

2. The centrifugal compressor (100) according to claim 1, wherein a blade array (2, 3) is provided upstream of the mixing section (10), the shape of the blade array being designed to impart a predetermined inclination to the lateral flow (30).

3. The centrifugal compressor (100) according to claim 1 further comprises a bladed distribution channel (9) located downstream of the mixing section (10), the distribution channel being passed through by both the main flow (20) and the side flow (30) and being configured to minimize the axial distance between the first impeller (7) and the second impeller (8) of the second stage, that is, the distance is no greater than the distance that can be obtained in a compressor without side flow and having a return flow channel, the volume flow processed by the return flow channel being equal to the volume flow generated by the mixing of the main flow (20) and the side flow (30) in the mixing section (10).

4. The centrifugal compressor (100) according to claim 1, wherein the nozzle (1) is tangential to the casing (14).

5. The centrifugal compressor (100) according to claim 1, wherein the cross section of the spiral volute (5) is such that the velocity (v) of the lateral flow (30) is constant and does not exceed 55 m / s.

6. The centrifugal compressor (100) of claim 2, wherein the blade array is an axial blade array (2).

7. The centrifugal compressor (100) according to claim 2, wherein the blade array is a radial blade array (3).

8. The centrifugal compressor (100) according to claim 2, wherein the blade array (2, 3) has a number of blades in the range of 20 to 50.

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

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