Multi-stage turbomachinery with lateral flow

Abstract

A multi-stage turbomachine (100) for processing a working fluid with at least one side flow (30) of the working fluid, the turbomachine having a casing (14), a shaft (16) supported by bearings (15), an inlet (12) for a main flow (20) of the working fluid, an outlet nozzle (17) for the main flow (20), and first and second impellers (7) and (8) cantilevered relative to the bearings (15). The turbomachine also has a nozzle (1) for introducing the side flow (30), a spiral scroll (5) housed within the casing (14) for imparting a degree of vorticity to the working fluid in the side flow (30), an annular channel (13) for transporting the side flow (30), and a mixing section (10) where the main flow (20) and the side flow (30) meet and mix.

Description

[Technical Field]

[0001] The present invention relates to multi-stage turbomachines, and in particular to a gas injection system for a centrifugal compressor having an impeller cantilevered relative to a shaft bearing. Although the following refers specifically to a multi-stage centrifugal compressor, the invention is equally applicable to any multi-stage turbomachine for processing a working fluid. [Background technology]

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

[0003] Centrifugal compressors may 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 precisely, each stage of a centrifugal compressor usually consists of a duct that sucks in the gas to be compressed, an impeller capable of providing kinetic energy to the gas, and a duct that connects the impeller to the next stage, whose task is to convert the kinetic energy of the gas coming from the impeller into pressure energy.

[0005] In particular, these ducts downstream of the impeller consist of a first section of duct exiting the impeller (called the diffuser), a substantially U-shaped junction (called the crossover), and a second section of duct flowing into the next impeller (called the return channel).

[0006] Multi-stage compressors are also characterized by the relative positions of the bearings for the various compression stages: in particular, the bearings are positioned outside the centrifugal stages in a so-called "between-bearing" configuration, or the centrifugal stages are positioned 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 in the intermediate stages (also called lateral flow). Some typical applications of these compressors are represented by machines used in heat pumps and refrigerator cycles, which use high molecular weight gases injected in the intermediate stages according to the process needs. Gas injection is usually performed using a plenum or volute, which is obtained in the stator section of the compressor, between two consecutive stages, in connection with external flanges. Compressors with such gas injection are also called "lateral flow compressors".

[0008] Figure 1 shows an example of a compressor with lateral flow ("injection inlet" and "injection scroll") and a "bearing-to-bearing" impeller configuration, and Figure 2a shows an example of a compressor with lateral flow ("injection scroll") and a "cantilever" configuration of the impeller.

[0009] Cantilever compressors can be coupled with a gearbox, i.e., a rotational speed increaser, and are called integral gear compressors (or IGCs, an acronym for "Integral Gear Compressor" in English). Figure 2b shows, for example, a three-stage compressor. A compressor with an integral speed increaser is a multi-shaft machine, with the shaft carrying the compressor impeller having pinion teeth formed on the same shaft. The shaft and associated pinion are arranged around a central gear connected to the primary shaft, which is in turn connected to the drive machine that drives the compressor. This type of compressor allows for a perfect combination of shaft speed and impeller size for each secondary shaft. All impellers are cantilevered and can be equipped with multiple adjustable blades at the inlet to guide the inlet flow. Interstage cooling of the gas flow can be performed after each discharge from the impeller.

[0010] The combination of these features enables a compact design with excellent energy efficiency, even at high flow rates and partial loads.

[0011] Among the limitations imposed by the IGC architecture, the compactness requirement limits the possibility of using large lateral flows in this configuration.

[0012] IGC architectures, especially those with more than two impellers, require the following: Limiting the axial length of the stage to keep the cantilever mass relatively close to the bearing. Excessive distance between the impeller and the bearing makes the compressor impractical from a rotordynamics standpoint, with vibrations that exceed acceptable levels. The presence of multiple compressors also limits the radial dimensions of the compressor casing, i.e., the size of the central gear wheel. In fact, the diameter of the central gear wheel determines the distance between the satellite gear pinions and therefore the distance between the axles of the various compressors installed. To ensure that the outer volutes of these compressors do not collide with each other, it is wise to use a central gear wheel with a large diameter, but the size of this wheel is limited by centrifugal forces acting as a function of diameter and speed.

[0013] In current state-of-the-art IGC compressor technology, at least one side flow is desired between two impellers mounted on the same pinion, making it difficult to design multiple compressors connected to the same central gear. This is due to design constraints on the dimensions of the central gear wheel (as previously mentioned) and the dimensions of side flow injection systems available according to existing technology. Figure 3 shows the current state of the art for side flow injection system geometries in both axial and radial size.

[0014] The so-called downstream injection geometry (downstream means of the return channel) makes it possible to increase the axial distance between the impeller and the bearing (FIG. 3a).

[0015] The so-called "upstream injection geometry" limits the axial distance but places the injection scroll on the outer diameter of the compressor, thereby increasing its radial size (Figure 3b).

[0016] Therefore, there is a need for a design approach for lateral flow injection systems that overcomes or at least mitigates the above-mentioned drawbacks. Summary of the Invention [Problem to be solved by the invention]

[0017] The present invention provides an improved side flow injection system in a multi-stage turbomachine having one or more side flows of working fluid. [Means for solving the problem]

[0018] The present invention relates particularly, although not exclusively, to centrifugal compressors in which the impeller is cantilevered from a bearing which supports the rotating shaft of the compressor.

[0019] More specifically, the present invention is suitable for architectures with an arrangement of several centrifugal compressors connected to a single rotational speed increaser according to the IGC (Integrated Gear Compressor) principle, i.e., where there is a central gear wheel driving a series of planetary pinions and therefore the compressors connected to them.

[0020] The present invention aims to overcome the current limitations of compressor side flow injection technology and enable the design of integrally geared compressors where more than two stages are housed around a common gear and with large intermediate injection flows between the impellers.

[0021] In particular, the new lateral flow configuration provides, upstream of the first impeller, an injection screw and a suitable channel (bladed in some implementations, compartmentless in others) for connecting the injection screw to a mixing section ("upstream injection") located upstream of the return channel. The channel connecting the injection screw to the mixing section (i.e., distributing the fluid circumferentially) is positioned at a greater radial distance and further outward ("crossover") than the connecting curve between the diffuser and the return channel.

[0022] The injection screw location increases the overall axial length of the compressor, but because it is mounted "upstream" of the first impeller, i.e., at the end of the shaft, it does not affect the overhang between the end impeller and the bearing.

[0023] Because the injection screw is positioned as close as possible to the compressor's axis of rotation, this component does not affect the overall radial dimension of the compressor compared to applications without side flow injection. The only slight increase in the overall radial dimension occurs due to the presence of a channel connecting the screw to the mixing section. This channel has a reduced size, typically in the range of 10-25% of the radial size, compared to screws positioned according to known techniques, due to the annular surface that extends across the entire outer diameter at the "crossover."

[0024] Advantageously, an array of blades is provided upstream of the mixing section, the blades being shaped to control the tangential velocity and slope of the lateral flow with minimum incidence losses at the entrance to the blade distribution channel.

[0025] According to one aspect of the present invention, a multi-stage turbomachine for processing a working fluid is described, provided with at least one lateral flow and having the features set out in the independent claims attached hereto.

[0026] Further preferred and / or particularly advantageous ways of implementing the aforementioned system are described according to the features set out in the attached dependent claims. [Brief explanation of the drawings]

[0027] The present invention will now be described with reference to the accompanying drawings, which show some non-limiting examples of implementations of the invention. [Figure 1] 1 shows a schematic representation of a compressor with a side flow and impeller between bearings configuration according to the known art; [Figure 2a] 1 shows a schematic diagram of a compressor with a side flow and cantilevered impeller configuration according to the known art; [Figure 2b] 1 shows a cross-sectional view of a compressor with an integral rotation multiplier according to the known art; [Figure 3] 1A-1C show schematic diagrams of two known configurations of a lateral flow injection system according to the known art; [Figure 4] 1 illustrates a schematic diagram of a compressor with a lateral flow and cantilevered impeller configuration according to a preferred embodiment of the present invention. [Figure 5] 5 shows a schematic diagram of the compressor of FIG. 4 in detail; DETAILED DESCRIPTION OF THE INVENTION

[0028] Referring to Figure 4, a multi-stage turbomachine 100 according to a preferred embodiment of the present invention will be described. This is a centrifugal compressor with a cantilevered impeller and lateral flow, and the same reference numeral 100 will be used for this centrifugal compressor. Figure 4 partially illustrates the compressor by omitting components that are not part of the present invention.

[0029] The proposed compressor 100 is a two-stage compressor with a side flow between the first and second stages. The following description of the design of components that allows for side flow injection within a compressor should also be considered valid for multi-stage compressors, and more generally for multi-stage turbomachines, since side flow injection can occur between two consecutive stages of a compressor and / or turbomachine.

[0030] The compressor 100 is provided with a casing 14 and a rotor assembly including a first impeller 7 associated with the first compression stage and a second impeller 8 associated with the second compression stage. A compressor shaft 16 (partially shown), rotatable about a rotation axis X, is supported by a bearing pair 15 (only one of which is shown in the figure, the other being positioned to the right of the illustrated bearing), and both the first impeller 7 and the second impeller 8 are cantilevered relative to the bearing pair 15. The compressor is configured to process a main flow 20 of working fluid which enters the compressor axially via an inlet 12 and, after being compressed in the first impeller 7 and then the second impeller 8, leaves the compressor 100 radially via an exhaust duct 17.

[0031] The new design involves lateral flow injection of working fluid 30 using an injection scroll positioned as close as possible to the compressor axis of rotation so as not to affect the compressor radial dimensions or the cantilevered impeller against the bearings.

[0032] The side flow 30 of the compressor 100, for example the flow coming from the separator or evaporator at a pressure level intermediate between the compressor suction and discharge, is fed to the compressor where it is mixed with the main flow from the first stage. The design must be such as to obtain low fluid pressure losses in the various ducts, for example 0.5 to 3% of the total pressure loss, while maintaining a minimum radial size, i.e., a minimum rotor projection (i.e., the distance of the first impeller from the bearing), without affecting the design of the speed increaser, thus having a robust design from a rotor dynamics point of view.

[0033] The new design is shown in Figures 4 and 5, where the lateral flow 30 comes from a nozzle 1 provided with a flange 11 and is introduced into the compressor 100. The nozzle 1 is preferably tangential to the casing 14 of the compressor 100 (see Figure 5) and follows the geometry of the subsequent components without causing excessive deviations in the gas flow. The lateral flow rate from the nozzle is distributed within the spiral scroll 5, which feeds the channels 13 evenly and imparts a certain degree of vorticity to the gas.

[0034] A feature of this new design compared to known technology is that the scroll 5 is located upstream of the diffuser 6 of the first impeller 7, and the size of the scroll does not affect the rotor design, e.g., the position of the impeller relative to the bearings.

[0035] The space between the diffuser 6 and the inlet 12 of the compressor 100 allows the scroll 5 to be sized with a large cross-sectional area, to have a lower velocity than typically achieved in known embodiments, preferably 55 m / s or less, e.g., 30-55 m / s. This lower velocity has a positive impact on fluid losses, which are proportional to the gas density times the square of its velocity. For applications such as heat pumps, the molecular weight of the fluid is preferably between 44 and 130 g mol -1 Since the density is directly proportional to the molecular weight of the gas, it is important to minimize the flow velocity.

[0036] The scroll 5 has a spiral shape with a cross section that gradually decreases in the circumferential direction, maintaining a constant gas velocity and thus a uniform supply of the channel 13. The scroll shape is such that it generates a circumferential movement of the gas around the rotation axis of the compressor 100. This vorticity of the gas that generates the lateral flow is necessary, as it follows a behavior similar to that of the gas coming 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 that is converted into pressure by the return channel.

[0037] The scrolls deploy inside the casing 14 of the compressor 100 and do not affect the radial dimension of the compressor, unlike what occurs in known embodiments (e.g., as shown in Figure 3b) where the scrolls affect the radial dimension of the compressor.

[0038] The scroll 5 is located at a radially outer position and feeds into an annular channel 13 adjacent to a U-shaped, small radially extending elbow 4. This approach minimizes the radial size of the entire compressor 100, typically in the range of 10-25% of the radial size of the screw positioned according to known techniques. The channel 13 is necessary to direct the side flow from the scroll 5 into the mixing section 10, where it begins to mix with the main flow coming from the first impeller 7 (first stage impeller) through the diffuser 6 and elbow 4.

[0039] Optimal hydrodynamic design for mixing two flow rates requires matching both the velocity and direction of the gases. Any difference in velocity or angle introduces an exchange of momentum, with an accompanying increase in pressure loss, as in any real conversion.

[0040] Considering a typical centrifugal compressor for heat pump and refrigeration duties, the main flow in the mixing section 10 typically has a velocity of 15-50 m / s and forms an angle of 15°-30° with respect to the tangential direction. Ideally, the side flows in the mixing zone 10 should have the same velocity and inclination values ​​to have uniform flow mixing with minimal losses.

[0041] The velocity and angle of the lateral flow in the mixing zone 10 depend on the annular area of ​​the channel 13 and the tangential vortex generated inside the scroll 5. In particular, the velocity in the channel 13 is inversely proportional to the cross-sectional area of ​​the channel, while the flow angle associated with the vortex follows the well-known behavior of a "free vortex" with constant gas momentum, which is the product of the tangential velocity and the distance of the cross section from the axis of rotation.

[0042] Since the new design goal is to minimize the velocity in the screw and thus the fluid loss, the vortices acting on the side flows can generate a side flow with a different inclination angle than the main flow in the mixing section 10. In other words, this means that the side flow direction moves further away from the tangential direction. This angle difference, typically 5° to 15°, cannot be compensated for by changing the area of ​​the channel 13. This is because the height of the channel 13 is set to control the adaptation of the velocity of the side flow relative to the main flow. In fact, it is known that the flow angle is given by the vector composition of the tangential and perpendicular components of the velocity, i.e., it characterizes the volumetric flow rate in the channel.

[0043] When this situation occurs, the solution envisaged by the present invention is to install a blade array upstream of the mixing section 10, which will provide the required gas deflection. As shown 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 in the range of 20 to 50, providing the required "solidity" of the blade array (defined as the ratio of the axial length to the relative circumferential pitch of the blade array).

[0044] The blade arrays (both axial blade array 2 and radial blade array 3) can be integrated into the body of the channel 13 (e.g., fabricated by co-casting with the diaphragm containing the blades), or can be assembled separately or machined into separate disks that are bolted to the body of the channel 13. The shape of the blades is typically a circular arc, but more complex geometries with three-dimensional airfoil cross sections can also be fabricated.

[0045] After mixing in the mixing section 10, both the main flow and the lateral flow are directed into the bladed distribution channel 9, which creates only a small 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 relative to the bearing 15 is another advantage compared to implementations according to known techniques (e.g., those in Figure 3a). In fact, the reduced overhang improves the rotor's dynamic behavior, allowing the use of rotors with shafts having smaller diameters. The reduction in shaft diameter has a significant positive impact on performance. This is because, with a smaller shaft, the flow entering the impeller has a lower relative velocity 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, which is proportional to the shaft diameter.

[0046] 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 (e.g., that of FIG. 3a). This allows for a more uniform flow to the second impeller 8, which has beneficial effects on the aerodynamics of the impeller. For example, due to the flow with minimal fluctuations in incidence at the inlet of its blades, and due to the lower intensity of flow disturbances, such as wakes and turbulence, which can adversely affect the performance of the second stage stator components up to the discharge nozzle 17, for example, impairing their ability to convert pressure into the dynamic components of the flow.

[0047] It should be understood that in addition to the methods for implementing the present invention as described above, there are numerous further variations. It should also be understood that the above implementation methods are merely exemplary and do not limit the scope of the invention, its applications, or possible configurations. Rather, while the above description will enable one skilled in the art to implement the invention according to at least one exemplary configuration, it should be understood that numerous variations of the described components are possible without departing from the scope of the invention as defined in the appended claims.

Claims

1. A centrifugal compressor (100) for processing a working fluid, comprising at least one side flow (30) of the working fluid, a casing (14); a shaft (16) rotating about an axis of rotation (X) and supported by bearings (15); an inlet (12) for the main flow (20) of working fluid; - an outlet nozzle (17) of the main flow (20); a rotor group including at least a first impeller (7) and at least a second impeller (8), both of which are cantilevered and on the same side of the bearing (15); The compressor comprises the following components to handle the lateral flow (30) of working fluid: a nozzle (1) for providing a lateral flow (30); a spiral scroll (5) for imparting a degree of vorticity to the working fluid in the lateral flow (30), the spiral scroll (5) being housed in a casing (14) upstream of the diffuser (6) of the first impeller (7); an annular channel (13) for carrying a lateral flow (30), located at a radially outer position, adjacent to the elbow (4), U-shaped and crossed by the main flow (20); - further comprising a mixing section (10) where the main flow (20) and the side flow (30) meet and mix with each other; The compressor (100) is characterized in that the components are configured so as not to affect the distance between the first impeller (7) and the second impeller (8) relative to the corresponding bearings (15) or to affect the radial dimensions of the compressor (100).

2. 2. The compressor (100) of claim 1, wherein a blade array (2, 3) is disposed upstream of the mixing section (10), the array being shaped to provide a predetermined gradient of the lateral flow (30).

3. 2. The compressor (100) of claim 1, further comprising a bladed distribution channel (9) downstream of the mixing section (10) where the main flow (20) and the side flow (30) intersect, and configured to minimize the axial distance between the first impeller (7) and the second impeller (8) of the second stage, i.e., to be equal to or less than that obtained in a compressor without a side flow, with a return channel handling a volumetric flow rate equal to the volumetric flow rate resulting from the mixing of the main flow (20) and the side flow (30) in the mixing section (10).

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

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

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

7. The compressor (100) of claim 2, wherein the blade array is an array of radial blades (3).

8. The compressor (100) of claim 2, wherein the blade array (2, 3) has a blade count of between 20 and 50.

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

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