MULTI-STAGE IMPELLER USABLE WITH A COMPRESSOR AND COMPRESSOR WITH MULTI-STAGE IMPELLER

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application claims the benefit of U.S. Provisional Application No. 63/426,545 filed on Nov. 18, 2023, titled: MULTI-STAGE IMPELLER USABLE WITH A COMPRESSOR AND COMPRESSOR WITH MULTI-STAGE IMPELLER, the entirety of which is incorporated by reference herein.

INTRODUCTION

The present disclosure generally relates to compressors, particularly to a compressor with a multi-stage impeller to improve the efficiency of a vapor compression cycle.

BACKGROUND

A multi-stage compressor system is known to improve the efficiency of a vapor compression cycle; however, multi-stage compressor systems that utilize multiple compressors in series have the disadvantage of increased complexity and maintenance. Further, the increased size and complexity of multi-stage compressor systems causes challenges with packaging the overall system.

A single-stage compressor system, specifically a single-stage centrifugal compressor system, is generally more compact and reliable due to decreased complexity.

One attempt at providing improved efficiency while decreasing complexity includes a compressor utilizing flash gas injected into a space normally defined by the compressor casing at either front or the rear shroud of the impeller. This system makes at least partial use of the disc friction in the compressor in order to impart kinetic energy to the flashed gas and to reduce the relative velocity of such gas to the velocity of gas leaving the impeller so as to minimize gas mixing losses. This system allows recovery of energy from flash gas, but the drawbacks of this system are that the compressor efficiency is generally decreased due to the impact of the economizer flow on the primary flow, and that cycle efficiency is cannot be optimized to the ideal flash pressure.

In recognizing the aforementioned deficiencies, the disclosure herein describes improvements to compressor and cycle efficiency while maintaining the simplicity and size advantages of a single-stage compressor.

SUMMARY

The following presents a simplified summary of one or more aspects of the invention to provide a basic understanding of such aspects. This summary is not an extensive overview of all contemplated aspects and is intended to neither identify key or critical elements of all aspects nor delineate the scope of any or all aspects. Its purpose is to present some concepts of one or more aspects in a simplified form as a prelude to the more detailed description that is presented later.

In one aspect, a compressor system includes a rotatable impeller configured to increase a pressure of one or more fluids from a compressor inlet to a compressor outlet. The compressor further includes a primary impeller section having a first set of impeller blades configured to pressurize a first fluid source; and a secondary impeller section with a second set of impeller blades configured to pressurize a second fluid source. In some aspects of the disclosure described herein, the first fluid source may be refrigerant received from an evaporator and the second fluid source may be a refrigerant received from an economizer and/or an intercooler.

In another aspect of this disclosure described herein, a refrigeration system includes a condenser, an evaporator, an expansion apparatus, an economizer or intercooler, and a compressor. The compressor includes a rotatable impeller with a primary impeller section having a first set of impeller blades and configured to pressurize a refrigerant received from the evaporator, and a secondary impeller section with a second set of impeller blades configured to pressurize refrigerant received from the economizer or intercooler.

In another aspect of the disclosure described herein, an impeller usable with a centrifugal compressor includes an impeller hub with a shrouded primary impeller section having a first set of impeller blades and configured to pressurize a first fluid source and a secondary impeller section with a second set of impeller blades connected to the shrouded primary impeller section and configured to pressurize a second fluid source.

These and other aspects of the invention will become more fully understood upon a review of the detailed description, which follows.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic diagram of one example of a refrigeration cycle usable with the compressor according to an aspect of the disclosure.

FIG. 2 is a partial cross-sectional view of an example compressor according to aspects of the disclosure.

FIG. 3 is a close-up cross-sectional view of a primary impeller and secondary impeller of a compressor according to one aspect of the disclosure.

FIG. 4 is a close-up cross-sectional view of a primary impeller and secondary impeller according to another aspect of the disclosure.

FIG. 5 is a partial isometric view of a secondary impeller according to one aspect of the disclosure.

FIG. 6 is a graph of pressure enthalpy characteristics, specifically log of pressure versus enthalpy, of a refrigeration cycle with a single compressor.

FIG. 7 is a graphic of pressure enthalpy characteristics, specifically log of pressure versus enthalpy, of a refrigeration cycle with a single compressor having a primary impeller and a secondary impeller according to an aspect of the disclosure.

FIG. 8 is a close-up cross-sectional view of a primary impeller and secondary impeller according to another aspect of the disclosure.

FIG. 9 is a close-up cross-sectional view of a primary impeller and secondary impeller according to another aspect of the disclosure.

FIG. 10A is partial view of an example of an impeller with a with a 45-degree (45°) exit angle.

FIG. 10B is a partial view of an example of an impeller with a with a 30-degree (30°) exit angle.

FIG. 10C is a partial view of an example impeller inlet blade profile usable with the impeller of FIG. 10A or 10B.

FIG. 10D is a partial view of an example impeller inlet blade profile usable with the impeller of FIG. 10A or 10B.

DETAILED DESCRIPTION

The detailed description set forth below in connection with the appended drawings is intended as a description of various configurations and is not intended to represent the only configurations in which the concepts described herein may be practiced. The detailed description includes specific details for the purpose of providing a thorough understanding of various concepts. However, it will be apparent to those skilled in the art that these concepts may be practiced without these specific details. In some instances, well known components are shown in block diagram form to avoid obscuring such concepts.

I. TERMINOLOGY

Throughout the disclosure, the terms substantially or approximately may be used as a modifier for a geometric relationship between elements or for the shape of an element or component. While the terms substantially or approximately are not limited to a specific variation and may cover any variation that is understood by one of ordinary skill in the art to be an acceptable variation, some examples are provided as follows. In one example, the term substantially or approximately may include a variation of less than 10% of the dimension of the object or component. In another example, the term substantially or approximately may include a variation of less than 5% of the object or component. If the term substantially or approximately is used to define the angular relationship of one element to another element, one non-limiting example of the term substantially or approximately may include a variation of 5 degrees or less. These examples are not intended to be limiting and may be increased or decreased based on the understanding of acceptable limits to one of skill in the relevant art.

For purposes of the disclosure, directional terms are expressed generally with relation to a standard frame of reference when the system and apparatus described herein is installed in an in-use orientation.

Terms such as a, an, and the are not intended to refer to only a singular entity, but include the general class of which a specific example may be used for illustration. The terms a, an, and the may be used interchangeably with the term at least one. The phrases at least one of and comprises at least one of followed by a list refers to any one of the items in the list and any combination of two or more items in the list. All numerical ranges are inclusive of their endpoints and non-integral values between the endpoints unless otherwise stated.

The terms first, second, third, and fourth, among other numeric values, may be used in this disclosure. It will be understood that, unless otherwise noted, those terms are used in their relative sense only. In particular, in some aspects certain components may be present in interchangeable and/or identical multiples (e.g., pairs). For these components, the designation of first, second, third, and/or fourth may be applied to the components merely as a matter of convenience in the description of one or more of the aspects of the disclosure.

The term fluid or fluid flow is used throughout the disclosure. While these terms are not limited to a specific type of fluid may cover any known fluid in the art, some examples include but are not limited to thermal fluids such as liquids, gasses, and vapors.

For context, an overview is provided of aspects of the disclosure and the advantages the disclosure provides. This overview, and the detailed description that follows, has been presented for purposes of illustration and description. It is not intended to be exhaustive nor to limit the disclosure to the forms described. Numerous modifications are possible considering the teachings herein, including a combination of the described aspects. Some potential modifications have been discussed and others will be understood by those skilled in the art. The various aspects were chosen and described to best illustrate the principles of the present disclosure and various aspects as are suited to the particular use contemplated. The scope of the present disclosure is, of course, not limited to the examples or aspects set forth herein but can be employed in any number of applications and equivalent devices by those of ordinary skill in the art. Rather, it is hereby intended the scope be defined by the claims appended hereto.

II. OVERVIEW

As described in further detail below, aspects of the disclosure described herein relate to an improved compressor system for improving efficiency of a refrigeration cycle while maintaining the simplicity and size advantages of a single-stage compressor. For example, a compressor system is disclosed that includes a single impeller that is configured to compress a fluid, such as refrigerant, when a rotational force is provided to the impeller. The impeller includes a first set of impeller blades as a primary impeller section and a secondary set of impeller blades as a secondary impeller section. The primary impeller section is configured to draw from and compress a first fluid source, such as refrigerant from an evaporator. The secondary impeller section is configured to draw from, and compress fluid from a secondary fluid source, such as refrigerant from an economizer and/or an intercooler. By integrating two or more impeller sections into a single impeller, the advantages of multi-stage compression in a thermodynamic cycle (e.g., a refrigeration cycle) can be utilized without the size, weight, and/or packaging disadvantages of using multiple compressors or multiple rotating impellers. In addition, a single rotating impeller allows for the aforementioned increases in efficiency of the cycle with the decreased maintenance of a single rotating impeller.

For context, a simplified diagram of a refrigeration system and cycle usable with the compressor system described herein is shown in FIG. 1. It is noted that the refrigeration cycle described with respect to FIG. 1 is not intended to be limiting and as such, aspects of the current disclosure are usable with and relevant to any thermodynamic cycle requiring the use of a compressor.

Referring to FIG. 1, an example refrigeration system 100 includes a condenser 102 for removing heating from refrigerant vapor until it condenses into a saturated liquid state, an evaporator 104 for absorbing heat from an environment by converting liquid refrigerant to a vapor, a compressor 110 for drawing or sucking refrigerant and compressing the refrigerant. The condenser 102 may be any type of condenser configured to reject heat from the refrigerant via the atmosphere and/or to condense water 103 flowing therethrough. The condenser 102 may receive refrigerant from the discharge line 108 of the compressor 110. Flow of the refrigerant through the condenser 102 may cause a phase change to liquid as heat is rejected from the refrigerant via the condenser 102. The cycle or system 100 may further include an economizer 106 configured to receive a liquid or substantially liquid refrigerant from the condenser 102 so that part of the total refrigerant flow from the condenser 102 cools the rest of the refrigerant flow and/or for the management of flash gas. While an economizer 106 is provided as an example, the economizer 106 may be used with or substituted with any known intercooler, which may be a surface or flash type. For example, the economizer 106 and/or intercooler may for example separate liquid refrigerant from any refrigerant that has boiled and/or has converted back to vapor, which may be referred to throughout the disclosure as flash gas or flash vapor. The aforementioned economizer 106 and/or intercooler may receive refrigerant from the condenser 102 via an intermediate expansion valve 111 in the fluid path between condenser 102 and the economizer 106.

The liquid refrigerant from the economizer 106 may then be provided to the expansion valve 109, causing the liquid refrigerant to expand and decrease in pressure and temperature. The liquid and gas refrigerant may then be provided to the evaporator 104 via a liquid line 114, wherein the expanded liquid is evaporated by absorbing heat from a fluid provided through a heat exchanger. In one example, the heat may be absorbed from a cycled water or other fluid source 105 that is cycled through the evaporator 104. The gaseous refrigerant from the evaporator 104 is then provided to the compressor 110 via a suction line 107, wherein the pressure and temperature of the gaseous refrigerant is increased when power is supplied to the compressor. As described in further detail below, the compressor 110 may include, for example, a plurality of impellers. For example, the compressor 110 may include a primary impeller 113 and a secondary impeller 115. The combination of the primary impeller 113 and secondary impeller 115 may be interchangeably referred to throughout the disclosure generally as an impeller. Further, the primary impeller 113 and the secondary impeller 115 may be referred interchangeably throughout this disclosure as a first impeller section and a second impeller section, respectively. The primary impeller 113 may for example have an inlet in fluid communication with the suction line 107 from the evaporator 104. The secondary impeller 115 may have, for example, an inlet in fluid communication with the economizer 106 via a flash gas line 112. The outlet or discharge line 108 of the compressor 110 may receive compressed gaseous refrigerant from both the primary impeller 113 and the secondary impeller 115. It is noted that while a primary impeller 113 and secondary impeller 115 are shown, additional impellers may be added to further improve efficiency. In addition, the disclosed compressor 110 described below may be used, for example, in a multi-stage compressor system. For example, the refrigeration system 100 may include a first compressor with a primary impeller 113 and secondary impeller 115 (or additional impellers) and the system may include a second compressor with either a single impeller or with a primary impeller 113 and a secondary impeller 115 (or additional impellers). In one example, the refrigerant vapor pressure provided to the secondary impeller 115 is a pressure that is greater than the pressure provided to the primary impeller 113 from the suction line 107 and is less than or equal to the output pressure from the compressor 110 provided to the discharge line 108.

III. DETAILED EXAMPLES

Referring to FIG. 2, an example compressor 210 may be analogous with compressor 110 in FIG. 1. The compressor 210 may include an input shaft 216 configured to receive a rotational force. The input shaft 216 may be operatively connected to a geartrain or drivetrain 217. The drivetrain 217 may be operatively connected to an impeller 205 with a primary impeller 213 and a secondary impeller 215. The primary impeller 213 may compress fluid supplied at a primary impeller inlet 212. The primary impeller inlet 212 may for example have one or more electronically or hydraulically controlled vanes 214 for controlling the flow of fluid into the primary impeller 213. The rotating primary impeller 213 compresses the fluid received at the primary impeller inlet 212 and discharges the compressed fluid to an outlet path 220.

The secondary impeller 215 is operatively connected to the primary impeller 213. In some examples, the primary impeller 213 and the secondary impeller 215 may be joined or formed/machined as a unitary structure. In other aspects, the secondary impeller 215 may be connectable to or otherwise connected to the primary impeller 213 via one or more known mechanical fasteners or joining methods. In some examples, the primary impeller 213 and/or the secondary impeller(s) may be additively manufactured as two or more separate components that are joined together or as a single unitary component. The secondary impeller 215 may for example be configured to be in fluid communication with a flash gas or vapor source, such as from an intercooler and/or economizer (e.g., economizer 106 in FIG. 1). The flash gas or vapor source (e.g., economizer and/or intercooler) may be in fluid communication with an inner cavity 218 of the compressor 210. The secondary impeller 215 may compress the fluid provided to the inner cavity 218 and discharge the compressed fluid at a secondary impeller discharge outlet (see non-limiting detailed examples in FIGS. 3A and 3B below). The secondary impeller discharge outlet and outlet 220 may be configured to be in fluid communication so that the compressed fluid from both the primary impeller 213 and the secondary impeller 215 mix at the outlet 220.

Additional details of a primary impeller and secondary impeller that are usable with the aspects of the disclosure are described in further detail below with respect to FIGS. 3, 4, 8, and 9.

Referring to FIG. 3, an example compressor 310, including additional details of example aspects disclosed, may be analogous with compressor 210 shown in FIG. 2 and/or compressor 100 in FIG. 1. Compressor 300 includes a primary impeller 313 and a secondary impeller operatively connected to a drivetrain (not shown in FIG. 3, but such as drivetrain 217 in FIG. 2). The primary impeller 313 may have a plurality of vanes or blades 334 configured to compress fluid supplied at a primary impeller inlet 312. It is noted that, in FIG. 3, only a subset of blades and/or vanes 334 are provided with reference numbers to prevent obstruction of other features in the figure. Further, it is noted that the blade and/or vane structure is not limited to the structure shown in FIG. 3 and may include any number of blades and/or vanes in any configuration without departing from the scope of the disclosure. The primary impeller inlet 312 may for example have one or more electronically or hydraulically controlled vanes (hidden from view in FIG. 3, see e.g., reference 214 in FIG. 2) for controlling the flow of fluid into the primary impeller 313.

The vanes and/or blades 334 of the rotating primary impeller 313 are configured to compress the fluid sucked into the primary impeller inlet 312 and are configured discharge the compressed fluid at the primary impeller tip 332 into the outlet path 320. In one example, the primary impeller 313 may have a dynamic shroud 333 that is fixed to and configured to rotate with the primary impeller 313. In one example, the dynamic shroud 333 may be a unitary structure and formed and/or machined as a single component. In another example, the dynamic shroud 333 may be fastened to or otherwise connected to the blades and/or vanes 334 of the dynamic shroud 333. The primary impeller 313 may be sealed to prevent fluid leakage via one or more seals 336 configured to sealingly engage with a sealing surface, surfaces, or sealing features 337 of the primary impeller 313. In one aspect, the one or more seals may comprise one or more radial or circular labyrinth seal(s) or a like seal that provides a tortuous path to prevent leakage of fluid at the rotational interface of the primary impeller 313. In some examples, one or more of the seal(s) 336 and/or features 337 may prevent or resist passage of fluid while providing non-contact with an opposing surface or sealing features by controlling the passage of fluid through a variety of chambers within the seal by centrifugal motion and/or the formation of controlled fluid vortices.

The secondary impeller 315 may be operatively connected to the primary impeller 313. In some examples, the primary impeller 313 and the secondary impeller 315 may be joined or formed/machined as a unitary structure. In other aspects, the secondary impeller 315 may be connectable to or otherwise connected with the primary impeller 313 via one or more mechanical fasteners or joining methods. For example, the secondary impeller 315 may be mounted to or otherwise joined to the dynamic shroud 333 of the primary impeller 313.

The secondary impeller 315 may have a significantly reduced size, diameter, and/or volume compared to the primary impeller 313. For example, the outer or major diameter of the secondary impeller 315 may be between 70% and 85% of the primary impeller 313 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 315 may be between 40% and 95% of the main or primary impeller 313 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 313 may be between 40% and 85% of the main impeller 313 outer or major diameter.

In addition, the blades of the secondary impeller may have a blade profile with a higher degree of twist compared to the primary impeller. For example, the blades of the secondary impeller 315 may have a blade outlet angle 5 degrees (5°) to 30° lower than the primary impeller 313. For reference, a blade outlet angle of 90° angle would include blades that are purely radial at the outlet. The reference example 90° outlet angle would provide the highest lift capacity but tend to reduce flow capacity range of the impeller.

FIG. 10A is a partial view of an impeller 1015a with impeller blades 1034a with a 45° outlet angle. An Impeller with blades at 45° at the outlet may be usable with the secondary impeller(s) described herein and may provide a good tradeoff. In one aspect, an impeller with blades at 45° at the exit may be used as the primary impeller(s) described herein. Lower blade exit angles (i.e., lower than) 45° are generally not used in compressors due to a decreased lift capacity. However, in the case of a secondary impeller(s) (i.e., secondary impeller 315) described herein, a lower blade exit angle (e.g., lower than) 45° may be preferred in one aspect and may provide the required lift for the economizer cycle described herein.

FIG. 10B shows one example of an impeller with blades 1034b at an angle of 30° at the outlet. The 30° blade angle at the outlet is usable with aspects of the secondary impeller(s) described herein. As can be seen when comparing the 45° exit blade profile of FIG. 10A with the 30° exit blade profile of FIG. 10b, as the blade exit angle decreases the blades look more “spiraled” or “twisted.” Thus, in one example, the blade profile of the secondary impeller described herein (i.e., secondary impeller 315) may be more spiraled or twisted than the blade profile of the primary impeller (i.e., primary impeller 313).

As noted above, the aforementioned blade angle profiles are provided as non-limiting examples. As noted above, the blade outlet profile of the secondary impeller(s) described herein may be 5°-30° lower than the blade outlet profile of the primary impeller(s) described herein.

FIGS. 10C and 10D show a partial view of a first example impeller intel with blades 1014c having a first blade inlet profile and a second example impeller inlet 1015d with blades 1034d having a second blade inlet profile. In some examples the first blade inlet profile or secondary blade inlet profile are usable with either one of or any combination of the blade exit profiles shown in FIGS. 10A and 10b.

In addition, the secondary impeller blades and/or vanes may be adjusted accordingly to reduce the compression ratio of the economizer fluid flow through the compressor compared to the suction line fluid flow through the compressor. In one example, the secondary impeller blades and/or vanes may be adjusted and the outer diameter or major diameter of the secondary impeller 315 may be reduced as described above to reduce the compression ratio.

In one example, which is usable with aspects described in this disclosure, the economizer refrigerant flow rate through the secondary impeller 315 of the compressor may be half of or approximately half of the flow rate of the primary suction line flow through the primary impeller 313.

In one example, which is usable with aspects described in this disclosure, the economizer refrigerant flow through the secondary impeller 315 of the compressor may be half or approximately half of the omega (i.e., head factor) of the primary suction line flow through the primary impeller 313. For example, if the head factor is defined by the following equation:

Ω = Δ ⁢ Hi / SoundSpeed 2 Equation ⁢ 1

With ΔHi=isentropic compression enthalpy, the head factor of economizer refrigerant flow may be half of or approximately half of the head factor of the suction line refrigerant flow.

Utilizing a primary impeller 313 and a secondary impeller 315 as described throughout this disclosure allows for optimization of the geometries of the primary impeller 313 and/or the secondary impeller 315 to allow for an overall increased efficiency of a refrigeration cycle while reducing the impact of or any decrease in efficiency that the compression of the economizer fluid flow by the secondary impeller 315 has on the compression of the suction line fluid flow by the primary impeller 313. For example, since the vapor from the economizer or intercooler (e.g., economizer fluid flow from flash gas line 112 in FIG. 1) requires less kinetic energy than is needed to properly compress the gas received from the evaporator via the suction line (e.g., suction line 107 in FIG. 1), the secondary impeller 315 may be adjusted (e.g., in any one or any combination of diameter, size, volume, and/or blade/vane configuration) to reduce the impact on efficiency that compression of the economizer fluid flow has on the compression of the suction line fluid flow. In one example, a tip diameter (e.g., a distance from axis RR to tips 352) of the secondary impeller 315 may be chosen to adjust the compression ratio of suction line fluid flow and/or the economizer fluid and the economizer pressure. In some examples, the distance from RR or major diameter from one primary impeller tip 332 to an impeller tip on the opposite side of the primary impeller 313 may be greater than a distance from axis RR or major diameter from one secondary impeller tip 352 to an impeller tip on the opposite side of the secondary impeller 315.

As shown in FIG. 3, the secondary impeller 315 may have one or more blades and/or vanes 354. It is noted in FIG. 3 a reduced number of blades and/or vanes 354 are provided with reference numbers to prevent obstruction of the figure. Further, while a specific vane/blade configuration is shown in FIG. 3, it is noted that any vane and/or blade configuration may be utilized without departing from the scope of the disclosure. The vanes and/or blades 354 of the rotating secondary impeller 315 are configured to compress the fluid received from the economizer and/or intercooler (for example via flash gas line 112 in FIG. 1), which may be interchangeably referred to herein as economizer fluid flow. In one example, the economizer fluid flow may be provided to an inner cavity 318 of the compressor casing. As denoted by the example dashed lines with arrowheads representing flows in FIG. 3, the economizer fluid may be sucked into or otherwise be provided to a secondary impeller inlet 335, and the compressed economizer fluid flow fluid may exit the secondary impeller 315 at the secondary impeller tip or tips 352. In one example, the secondary impeller tip or tips 352 may be in fluid communication with the outlet path 320, thus causing the compressed economizer flow and suction line fluid flow to mix or partially mix before exiting the compressor 310. In one example, the compressor 310 may include a separation wall 371 at the outlet of the secondary impeller 315.

The separation wall 371 may be utilized to delay the mixing of the economizer fluid flow and the suction line fluid flow until the suction line fluid flow velocity has decreased. Delaying the mixing of the economizer fluid flow and the suction line fluid flow until the velocity difference between the suction line fluid flow and the economizer fluid flow has decreased reduces any mixing losses that may otherwise occur. Any one or any combination of the separation wall length, geometry, and/or the positional relationship of the separation wall with respect to the primary impeller tip(s) 332 and/or the secondary impeller tip(s) 352 may be optimized to decrease the velocity difference between the economizer fluid flow and the suction line fluid flow at the divergence of the compressed economizer fluid flow and the suction line fluid flow at the outlet path 320.

In one example, the secondary impeller 315 may also have a dynamic shroud 358 that is fixed to and configured to rotate with the secondary impeller 315. In one example, the dynamic shroud may be a unitary structure and formed and/or machined as a single component with the primary impeller 313 and the secondary impeller 315. In another example, the dynamic shroud 358 may be fastened to or otherwise connected to the blades and/or vanes 354 of the dynamic shroud 358. The secondary impeller 315 may be sealed to prevent fluid leakage via one or more seals 356 configured to sealingly engage with a sealing surface, surfaces, or sealing features 357 of the secondary impeller 315. In one aspect, the one or more seals may for example comprise one or more radial or circular labyrinth seal(s), which may be similar to or include one or any combination of features of the seal or seals 336 and/or the sealing surfaces or features 337 of the primary impeller 313 described in detail above. As mentioned above, the tip diameter of the secondary impeller (e.g., the major diameter of the secondary impeller at the compression side) may be adjusted or otherwise optimized to effect a desired compression ratio of the economizer fluid flow and/or the suction line fluid flow as described above.

Referring to FIG. 4, an example compressor 410, including additional details of example aspects disclosed and may be analogous with compressor 210 shown in FIG. 2 and/or compressor 100 in FIG. 1 and may further share features of the compressor 310 in FIG. 3. In one example, the primary difference between the compressor 310 and compressor 410 may be a floating stationary or fixed (i.e., non-rotating) shroud 458, whereas the compressor 310 in FIG. 3 utilizes a dynamic (i.e., rotating) shroud. The stationary shroud 458 may be biased or otherwise configured to have a resting net force in towards the blades and/or vanes 454 of the secondary impeller 415 (e.g., as indicated by arrow FF). The force FF may be provided by one or more springs and/or any biasing member or members. Since the fixed (floating) shroud 458 does not rotate with the blades and/or vanes of the secondary impeller 415, the clearance between the blade/vane tips and the shroud can be decreased as low as possible while preventing and/or decreasing friction between the blade/vane tips and the shroud 458. In one example implementation, the clearance between the surface of the shroud 458 facing the secondary impeller blade/vanes 454 during operation of compressor may be less than 50 micrometers (μm). In another example implementation, the clearance between the surface of the shroud 458 facing the secondary impeller blade/vanes 454 during operation of the compressor may be between 5 μm and 15 μm. In yet another implementation, the clearance may be between 7 μm and 13 μm or more preferably approximately 10 μm. In general, it is desirable to decrease the clearance as much as possible while reducing or eliminating friction due to contact between the blade/vanes 454 and the shroud 458.

In one aspect of the disclosure, the blades and/or vanes 454 of the secondary impeller 415 may be configured to provide a pressure field and/or aerodynamic thrust during rotation of the secondary impeller 415. The pressure field and/or aerodynamic thrust may counteract the biasing force FF causing the desired clearance between the blade/vane 454 tips and the shroud 458 during operation of the compressor 410. The pressure field loading capacity of the blades/vanes of the secondary impeller 415 and working fluid can be calculated by applying Navier-Stokes Equations for thin fluid film. The aforementioned Navier-Stokes Equations may be solved using Computerized Fluid Dynamics (CFD) software. The aforementioned calculation can be simplified to the Reynolds equation for a perpendicular rotating plane to an axis of rotation (thrust). For example, the following equation can be used to calculate the pressure field loading capacity: μ

∂ ∂ x ( ρ ⁢ h 3 12 ⁢ μ ⁢ ∂ P ∂ x ) + ∂ ∂ z ( ρ ⁢ h 3 12 ⁢ μ ⁢ ∂ P ∂ z ) = ∂ ∂ x ( ρ ⁢ Uh 2 ) + ∂ ∂ t ( ρ ⁢ h ) Equation ⁢ 2

or for cylindrical rotating surface (bearing) using the following equation:

∂ ∂ θ ( P _ ⁢ h _ 3 ⁢ ∂ P _ ∂ θ ) + ( R L ) 2 ⁢ ∂ ∂ z _ ( P _ ⁢ h _ 3 ⁢ ∂ P _ ∂ z _ ) = Λ ⁢ ∂ ∂ θ ( P _ ⁢ h _ ) + σ ⁢ ∂ ∂ t _ ( P _ ⁢ h _ ) , Equation ⁢ 3 Λ = 6 ⁢ μ ⁢ Ω ⁢ R 2 P a ⁢ C 2 , σ = 12 ⁢ μ ⁢ ω ⁢ R 2 P a ⁢ C 2

The variables of Equations 1 and 2 are outlined in Table 1 below:

TABLE 1
C	Jeu	Clearance
h	Epaisseur du film	film thickness
L	Longueur	Length
P	Pression du gaz	Gas pressure
Pa	Pression ambiante	Ambient pressure
R	Rayon	Radius
U, V	Vitesses linéaires d'une paroi	Linear velocities of a wall
X, Y, Z	Axes du systéme de coordonnées	Coordinate system axes
Ω	Vitesse de rotation	Rotation speed
μ	\fiscosité dynamique	dynamic viscosity
φ	Angle de calage	Attitude angle

Using the aforementioned Navier-Stokes Equations or example Reynolds Equations provided above, the geometry of the secondary impeller 415 and/or the blades/vanes 454 of the secondary impeller 415 can be optimized to achieve a desired clearance (e.g., the clearances described above) between the blades/vanes 454 and the shroud 458 while reducing or eliminating friction due to contact between the blades/vanes 454 and the shroud 458. Additional details of a blade/vane profile that may be relevant to the described secondary impeller 415 are described with respect to FIG. 5 below.

As mentioned above, the compressor 410 of FIG. 4 may have a primary impeller 413 and a secondary impeller 415 operatively connected to a drivetrain (e.g., drivetrain 217 in FIG. 2). The primary impeller 413 may have a plurality of vanes or blades 434 configured to compress fluid supplied at a primary impeller inlet 412. It is noted that, in FIG. 4, only a subset of blades and/or vanes 434 are provided with reference numbers to prevent obstruction of other features in the figure. Further, it is noted that the blade and/or vane structure is not limited to the structure shown in FIG. 4 and may include any number of blades and/or vanes in any configuration without departing from the scope of the disclosure. The primary impeller inlet 412 may for example have one or more electronically or hydraulically controlled vanes (hidden from view in FIG. 4, see, e.g., reference 214 in FIG. 2) for controlling the flow of fluid into the primary impeller 413.

The vanes and/or blades 434 of the rotating primary impeller 413 are configured to compress the fluid sucked into the primary impeller inlet 412 and are configured discharge the compressed fluid at the primary impeller tip 432 into the outlet path 420. In one example, the primary impeller 413 may have a dynamic shroud 433 that is fixed to and configured to rotate with the primary impeller 313. In one example, the dynamic shroud may be a unitary structure and formed and/or machined as a single component. In another example, the dynamic shroud 433 may be fastened to or otherwise connected to the blades and/or vanes 434 of the dynamic shroud 433. The primary impeller 413 may be sealed to prevent fluid leakage via one or more seals 436 configured to sealingly engage with a sealing surface, surfaces, or sealing features 437 of the primary impeller 413. In one aspect, the one or more seals may for example comprise one or more radial or circular labyrinth seal(s) or any type of seal that provides a tortuous path to prevent leakage of fluid at the rotational interface of the primary impeller. In some examples, one or more of the seal(s) 436 and/or features 437 may prevent or resist passage of fluid while providing non-contact with an opposing surface or sealing feature(s) by controlling the passage of fluid through a variety of chambers within the seal by centrifugal motion and/or the formation of controlled fluid vortices.

The secondary impeller 415 may be operatively connected to the primary impeller 413. In some examples, the primary impeller 413 and the secondary impeller 415 may be joined or formed/machined as a unitary structure. In other aspects, the secondary impeller 415 may be connectable to or otherwise connected with the primary impeller 413 via one or more mechanical fasteners or joining methods. For example, the secondary impeller 415 may be mounted to or otherwise joined to the dynamic shroud 433 of the primary impeller 413.

The secondary impeller 415 may have a significantly reduced size, diameter, and/or volume compared to the primary impeller 413. For example, the outer or major diameter of the secondary impeller 415 may be between 70% and 85% of the primary impeller 413 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 415 may be between 40% and 95% of the main or primary impeller 413 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 413 may be between 40% and 85% of the main impeller 413 outer or major diameter.

In addition, the blades of the secondary impeller 415 may have a blade profile with a higher degree of twist compared to the primary impeller. For example, the blades of the secondary impeller 415 may have a blade outlet angle 5 degrees (5°) to 30° lower than the primary impeller 413. For reference/comparison, a blade outlet angle of 90° angle would include blades that are purely radial at the outlet. The reference example 90° outlet angle would provide the highest lift capacity but tend to reduce flow capacity range of the impeller.

FIG. 10A is a partial view of an impeller 1015a with impeller blades 1034a with a 45° outlet angle. An Impeller with blades at 45° at the outlet may be usable with the secondary impeller(s) described herein and may provide a good tradeoff. In one aspect, an impeller with blades at 45° at the exit may be used as the primary impeller(s) described herein. Lower blade exit angles (i.e., lower than) 45° are generally not used in compressors due to a decreased lift capacity. However, in the case of a secondary impeller(s) described herein, a lower blade exit angle (e.g., lower than) 45° may be preferred in one aspect and may provide the required lift for the economizer cycle described herein.

FIG. 10B shows one example of an impeller with blades 1034b at an angle of 30° at the outlet. The 30° blade angle at the outlet is usable with aspects of the secondary impeller(s) described herein. As can be seen when comparing the 45° exit blade profile of FIG. 10A with the 30° exit blade profile of FIG. 10b, as the blade exit angle decreases the blades look more “spiraled” or “twisted.” Thus, in one example, the blade profile of the secondary impeller described herein may be more spiraled or twisted than the blade profile of the primary impeller.

As noted above, the aforementioned blade angle profiles are provided as non-limiting examples. As noted above, the blade outlet profile of the secondary impeller(s) described herein may be angled 5°-30° lower than the blade outlet profile of the primary impeller(s) described herein.

FIGS. 10C and 10D show a partial view of a first example impeller intel with blades 1014c having a first blade inlet profile and a second example impeller inlet 1015d with blades 1034d having a second blade inlet profile. In some examples the first blade inlet profile or secondary blade inlet profile are usable with either one of or any combination of the blade exit profiles shown in FIGS. 10A and 10b.

In addition, the secondary impeller blades and/or vanes may be adjusted accordingly to reduce the compression ratio of the economizer fluid flow through the compressor compared to the suction line fluid flow through the compressor. Utilizing a primary impeller 413 and a secondary impeller 415 as described throughout this disclosure allows for optimization of the geometries of the primary impeller 413 and/or the secondary impeller 415 to allow for an overall increased efficiency of a refrigeration cycle while reducing the impact of or any decrease in efficiency that the compression of the economizer fluid flow by the secondary impeller 415 has on the compression of the suction line fluid flow by the primary impeller 413. For example, since the vapor from the economizer or intercooler (e.g., economizer fluid flow from flash gas line 112 in FIG. 1) requires less kinetic energy than is needed to properly compress the gas received from the evaporator via the suction line (e.g., suction line 107 in FIG. 1), the secondary impeller may be adjusted (e.g., in any one or any combination of diameter, size, volume, and/or blade/vane configuration) to reduce the impact on efficiency that compression of the economizer fluid flow has on the compression of the suction line fluid flow.

In one example, a tip diameter of the secondary impeller 415 may be chosen to adjust the compression ratio of suction line fluid flow and/or the economizer fluid and the economizer pressure. In some examples, the distance from RR or major diameter from one primary impeller tip 432 to an impeller tip on the opposite side of the primary impeller 413 may be greater than a distance from axis RR or major diameter from one secondary impeller tip 452 to an impeller tip on the opposite side of the secondary impeller 415.

As shown in FIG. 4, the secondary impeller 415 may have one or more blades and/or vanes 454. It is noted that, in FIG. 4, only a subset of blades and/or vanes 454 are provided with reference numbers to prevent obstruction of other features in the figure. Further, while a specific vane/blade configuration is shown in the figure, it is noted that any vane and/or blade configuration may be utilized without departing from the scope of the disclosure. The vanes and/or blades 454 of the rotating secondary impeller 415 are configured to compress the fluid received from the economizer and/or intercooler (for example via flash gas line 112 in FIG. 1), which may be interchangeably referred to herein as economizer fluid flow. In one example, the economizer fluid flow may be provided to an inner cavity 418 of the compressor casing. As denoted by the example dashed lines with arrowheads representing flows in FIG. 4, the economizer fluid may be sucked into or otherwise be provided to a secondary impeller inlet 435, and the compressed economizer fluid flow fluid may exit the secondary impeller 415 at the secondary impeller outlet tip or tips 452.

In one example, the secondary impeller outlet tip or tips 452 may be in fluid communication with the outlet path 420 thus causing the compressed economizer flow and suction line fluid flow to mix or partially mix before exiting the compressor 410. In one example, the compressor 410 may include a separation wall 471 at the outlet of the secondary impeller 415. The separation wall 471 may be utilized to delay the mixing of the economizer fluid flow and the suction line fluid flow until the suction line fluid flow velocity has decreased. Delaying the mixing of the economizer fluid flow and the suction line fluid flow until the velocity difference between the suction line fluid flow and the economizer fluid flow has decreased reduces any mixing losses that may otherwise occur. Any one or any combination of the separation wall length, geometry, and/or the positional relationship of the separation wall with respect to the primary impeller tip(s) 432 and/or the secondary impeller tip(s) 452 may be optimized to decrease the velocity difference between the economizer fluid flow and the suction line fluid flow at the divergence of the compressed economizer fluid flow and the suction line fluid flow at the outlet path 420.

Referring to FIG. 5, an example profile of blades of an impeller 515 includes a specific configuration of surfaces usable with aspects of disclosure. In one example, the impeller 515 may be used in conjunction with the floating shroud 458 in FIG. 4 or in FIG. 8 described below. For example, the blades 544 of FIG. 5 may be analogous with the secondary impeller blades/vanes 454 of FIG. 4 or secondary impeller blades/vanes of 954 in FIG. 9. In one aspect of the disclosure, the blades 554 of the impeller 515 may include a narrowed blade section 554a with a narrower cross-section than a widened portion 554b closer to a tip of the blade 554. The tip of blade 554 may have a tip surface 554e. The tip surface 554e may be flat or substantially flat. In some aspects, the tip surface 554e may be formed to match or substantially match the geometry of a floating shroud surface that will be closest to the tip surface 554e so as to minimize the clearance therebetween, and thus improve a seal or reduce a pressure loss between the tip surface 554e and the floating shroud (e.g., floating shroud 458 in FIG. 4). As shown in FIG. 5, the profile of the blades 554 of the impeller 515 may further include a chamfer or curved portion 554c at the transition between the narrowed blade section 554a and the widened portion 554b. In addition, the blade 554 may be curved or chamfered at the edges of the tip surface 554e to reduce friction in case any contact occurs between the tip surface 554e and the floating shroud. In one aspect, the increase in cross-sectional area at the widened portion 554b and/or the curved or otherwise chamfered portion 554c may improve the ability of the impeller 515 to provide a pressure field and/or aerodynamic thrust during rotation of the secondary impeller. The pressure field and/or aerodynamic thrust may cause the shroud (e.g., shroud 458 in FIG. 4) to counteract a biasing force applied to the shroud (e.g., force FF in FIG. 4) and thus float to the desired clearance between tip surface 554e and the shroud (e.g. shroud 458 in FIG. 4) during operation of the compressor. As mentioned above, the pressure field loading capacity of the blades 554 of the impeller 515 and working fluid can be calculated by solving Navier-Stokes Equations. In an example implementation, the impeller 515 and the floating shroud (e.g., shroud 458 in FIG. 4) may be configured to provide a clearance during operation that is less than 50 micrometers (μm). In another example implementation, the impeller 515 and floating shroud may be configured to provide a clearance between 5 μm and 15 μm. In yet another example implementation, the aforementioned clearance (gap) may be between 7 μm and 13 μm or more preferably approximately 10 μm. In general, it is desirable to decrease the aforementioned clearance as much as possible while reducing or eliminating friction due to contact between the blades of the impeller 515 and the shroud.

It is noted that while several examples are provided above, a compressor or impeller in accordance with aspects of the disclosure may be varied for packaging or for retrofit (or ease of modifying a standard single-stage compressor). For example, the secondary impeller section described above instead may be closer to the hub or input shaft of the compressor while the primary impeller section is further from the input shaft of the compressor (i.e., a reversed configuration from the examples provided above).

Referring to FIGS. 6 and 7, example pressure enthalpy diagrams 600 and 700, respectively, of a single stage compressor (FIG. 6) and one example of a system in accordance with the current disclosure (FIG. 7).

Referring to FIG. 8, an example compressor 810, including additional details of example aspects disclosed, may be analogous with compressor 210 shown in FIG. 2 and/or compressor 100 in FIG. 1 and may further share features of the compressor 310 in FIG. 3, and/or compressor 410 in FIG. 4. Compressor 800 includes a primary impeller 813 and a secondary impeller operatively connected to a drivetrain (not shown in FIG. 3, but such as drivetrain 217 in FIG. 2). The primary impeller 813 may have a plurality of vanes or blades 834 configured to compress fluid supplied at a primary impeller inlet 812. It is noted that, in FIG. 3, only a subset of blades and/or vanes 834 are provided with reference numbers to prevent obstruction of other features in the figure. Further, it is noted that the blade and/or vane structure is not limited to the structure shown in FIG. 8 and may include any number of blades and/or vanes in any configuration without departing from the scope of the disclosure. The primary impeller inlet 812 may for example have one or more electronically or hydraulically controlled vanes (hidden from view in FIG. 8, see e.g., reference 214 in FIG. 2) for controlling the flow of fluid into the primary impeller 813.

The vanes and/or blades 834 of the rotating primary impeller 813 are configured to compress the fluid sucked into the primary impeller inlet 812 and are configured discharge the compressed fluid at the primary impeller tip 832 into the outlet path 820. In one example, the primary impeller 813 may have a dynamic shroud 833 that is fixed to and configured to rotate with the primary impeller 813. In one example, the dynamic shroud 833 may be a unitary structure and formed and/or machined as a single component. In another example, the dynamic shroud 833 may be fastened to or otherwise connected to the blades and/or vanes 834 of the dynamic shroud 833. The primary impeller 813 may be sealed to prevent fluid leakage via one or more seals 836 configured to sealingly engage with a sealing surface, surfaces, or sealing features 837 of the primary impeller 813. In one aspect, the one or more seals may comprise one or more radial or circular labyrinth seal(s) or a like seal that provides a tortuous path to prevent leakage of fluid at the rotational interface of the primary impeller 813. In some examples, one or more of the seal(s) 836 and/or features 837 may prevent or resist passage of fluid while providing non-contact with an opposing surface or sealing features by controlling the passage of fluid through a variety of chambers within the seal by centrifugal motion and/or the formation of controlled fluid vortices.

The secondary impeller 815 may be operatively connected to the primary impeller 813. One notable difference between the impeller of FIG. 3 and FIG. 8 is that the outer diameter or major diameter of the secondary impeller 815 is larger than the secondary impeller 313 of FIG. 3. In other words, the difference in diameters between primary impeller 813 and secondary impeller 815 is smaller than the difference in diameter between the primary impeller 313 and secondary impeller 315. In some examples, the primary impeller 813 and secondary impeller 815 may have the same diameter or major diameter. In one example, the number of veins, profile of the veins and/or height of the veins of the secondary impeller 315 may be modified to ensure that the proper pressure ratio or compression ratio is achieved. As noted above, the pressure ratio may be optimized using equations 1, 2, and 3 noted above. For example, the height of the impeller blades 854 of the secondary impeller 815 may be less that the height of the impeller blades 354 of the secondary impeller 315 in FIG. 3.

In some examples, the primary impeller 813 and the secondary impeller 315 may be joined or formed/machined as a unitary structure. In other aspects, the secondary impeller 815 may be connectable to or otherwise connected with the primary impeller 813 via one or more mechanical fasteners or joining methods. For example, the secondary impeller 815 may be mounted to or otherwise joined to the dynamic shroud 833 of the primary impeller 813.

The secondary impeller 815 may have a significantly reduced size, diameter, and/or volume compared to the primary impeller 813. For example, the outer or major diameter of the secondary impeller 815 may be between 70% and 85% of the primary impeller 813 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 815 may be between 40% and 95% of the main or primary impeller 813 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 813 may be between 40% and 85% of the main impeller 813 outer or major diameter.

In addition, the blades of the secondary impeller may have a blade profile with a higher degree of twist compared to the primary impeller. For example, the blades of the secondary impeller 815 may have a blade outlet angle 5 degrees (5°) to 30° lower than the primary impeller 813. For reference, a blade outlet angle of 90° angle would include blades that are purely radial at the outlet. The reference example 90° outlet angle would provide the highest lift capacity but tend to reduce flow capacity range of the impeller.

FIG. 10A is a partial view of an impeller 1015a with impeller blades 1034a with a 45° outlet angle. An Impeller with blades at 45° at the outlet may be usable with the secondary impeller(s) described herein and may provide a good tradeoff. In one aspect, an impeller with blades at 45° at the exit may be used as the primary impeller(s) described herein. Lower blade exit angles (i.e., lower than) 45° are generally not used in compressors due to a decreased lift capacity. However, in the case of a secondary impeller(s) described herein, a lower blade exit angle (e.g., lower than) 45° may be preferred in one aspect and may provide the required lift for the economizer cycle described herein.

FIG. 10B shows one example of an impeller with blades 1034b at an angle of 30° at the outlet. The 30° blade angle at the outlet is usable with aspects of the secondary impeller(s) described herein. As can be seen when comparing the 45° exit blade profile of FIG. 10A with the 30° exit blade profile of FIG. 10b, as the blade exit angle decreases the blades look more “spiraled” or “twisted.” Thus, in one example, the blade profile of the secondary impeller described herein may be more spiraled or twisted than the blade profile of the primary impeller.

As noted above, the aforementioned blade angle profiles are provided as non-limiting examples. As noted above, the blade outlet profile of the secondary impeller(s) described herein may be angled 5°-30° lower than the blade outlet profile of the primary impeller(s) described herein.

FIGS. 10C and 10D show a partial view of a first example impeller intel with blades 1014c having a first blade inlet profile and a second example impeller inlet 1015d with blades 1034d having a second blade inlet profile. In some examples the first blade inlet profile or secondary blade inlet profile are usable with either one of or any combination of the blade exit profiles shown in FIGS. 10A and 10b.

In addition, the secondary impeller blades and/or vanes may be adjusted accordingly to reduce the compression ratio of the economizer fluid flow through the compressor compared to the suction line fluid flow through the compressor. In one example, the secondary impeller blades and/or vanes may be adjusted as described above to reduce the compression ratio.

In one example, which is usable with aspects described in this disclosure, the economizer refrigerant flow rate through the secondary impeller of the compressor may be half of or approximately half of the flow rate of the primary suction line flow through the primary impeller 31.

In one example, which is usable with aspects described in this disclosure, the economizer refrigerant flow through the secondary impeller of the compressor may be half or approximately half of the omega (i.e., head factor) of the primary suction line flow through the primary impeller 31. For example, if the head factor is defined by Equation 1, as outlined above.

Utilizing a primary impeller 813 and a secondary impeller 815 as described throughout this disclosure allows for optimization of the geometries of the primary impeller 813 and/or the secondary impeller 815 to allow for an overall increased efficiency of a refrigeration cycle while reducing the impact of or any decrease in efficiency that the compression of the economizer fluid flow by the secondary impeller 815 has on the compression of the suction line fluid flow by the primary impeller 813. For example, since the vapor from the economizer or intercooler (e.g., economizer fluid flow from flash gas line 112 in FIG. 1) requires less kinetic energy than is needed to properly compress the gas received from the evaporator via the suction line (e.g., suction line 107 in FIG. 1), the secondary impeller 815 may be adjusted (e.g., in any one or any combination of diameter, size, volume, and/or blade/vane configuration) to reduce the impact on efficiency that compression of the economizer fluid flow has on the compression of the suction line fluid flow. In one example, a tip diameter (e.g., a distance from axis RR to tips 852) of the secondary impeller 815 may be chosen to adjust the compression ratio of suction line fluid flow and/or the economizer fluid and the economizer pressure. In some examples, the distance from RR or major diameter from one primary impeller tip 832 to an impeller tip on the opposite side of the primary impeller 313 may be greater than a distance from axis RR or major diameter from one secondary impeller tip 852 to an impeller tip on the opposite side of the secondary impeller 815.

As shown in FIG. 8, the secondary impeller 815 may have one or more blades and/or vanes 854. It is noted in FIG. 8, a reduced number of blades and/or vanes 854 are provided with reference numbers to prevent obstruction of the figure. Further, while a specific vane/blade configuration is shown in FIG. 8, it is noted that any vane and/or blade configuration may be utilized without departing from the scope of the disclosure. The vanes and/or blades 854 of the rotating secondary impeller 815 are configured to compress the fluid received from the economizer and/or intercooler (for example via flash gas line 112 in FIG. 1), which may be interchangeably referred to herein as economizer fluid flow. In one example, the economizer fluid flow may be provided to an inner cavity 818 of the compressor casing. As denoted by the example dashed lines with arrowheads representing flows in FIG. 8, the economizer fluid may be sucked into or otherwise be provided to a secondary impeller inlet 835, and the compressed economizer fluid flow fluid may exit the secondary impeller 815 at the secondary impeller tip or tips 852. In one example, the secondary impeller tip or tips 852 may be in fluid communication with the outlet path 820, thus causing the compressed economizer flow and suction line fluid flow to mix or partially mix before exiting the compressor 810. In one example, the compressor 810 may include a separation wall 871 at the outlet of the secondary impeller 815. The separation wall 871 may be utilized to delay the mixing of the economizer fluid flow and the suction line fluid flow until the suction line fluid flow velocity has decreased. Delaying the mixing of the economizer fluid flow and the suction line fluid flow until the velocity difference between the suction line fluid flow and the economizer fluid flow has decreased reduces any mixing losses that may otherwise occur. Any one or any combination of the separation wall length, geometry, and/or the positional relationship of the separation wall with respect to the primary impeller tip(s) 832 and/or the secondary impeller tip(s) 852 may be optimized to decrease the velocity difference between the economizer fluid flow and the suction line fluid flow at the divergence of the compressed economizer fluid flow and the suction line fluid flow at the outlet path 820.

In one example, the secondary impeller 815 may also have a dynamic shroud 858 that is fixed to and configured to rotate with the secondary impeller 815. In one example, the dynamic shroud may be a unitary structure and formed and/or machined as a single component with the primary impeller 813 and the secondary impeller 815. In another example, the dynamic shroud 858 may be fastened to or otherwise connected to the blades and/or vanes 854 of the dynamic shroud 858. The secondary impeller 815 may be sealed to prevent fluid leakage via one or more seals 856 configured to sealingly engage with a sealing surface, surfaces, or sealing features 857 of the secondary impeller 815. In one aspect, the one or more seals may for example comprise one or more radial or circular labyrinth seal(s), which may be similar to or include one or any combination of features of the seal or seals 836 and/or the sealing surfaces or features 837 of the primary impeller 813 described in detail above. As mentioned above, the tip diameter of the secondary impeller (e.g., the major diameter of the secondary impeller at the compression side) may be adjusted or otherwise optimized to effect a desired compression ratio of the economizer fluid flow and/or the suction line fluid flow.

Referring to FIG. 9, an example compressor 910, including additional details of example aspects disclosed and may be analogous with compressor 210 shown in FIG. 2 and/or compressor 100 in FIG. 1 and may further share features of the compressor 310 in FIG. 3, compressor 410 in FIG. 4 and/or compressor 810 in FIG. 8. In one example, the primary difference between the compressor 810 and compressor 910 may be a floating stationary or fixed (i.e., non-rotating) shroud 958, whereas the compressor 810 in FIG. 8 utilizes a dynamic (i.e., rotating) shroud. The stationary shroud 958 may be biased or otherwise configured to have a resting net force in towards the blades and/or vanes 954 of the secondary impeller 915 (e.g., as indicated by arrow FF). The force FF may be provided by one or more springs and/or any biasing member or members. Since the fixed (floating) shroud 958 does not rotate with the blades and/or vanes of the secondary impeller 915, the clearance between the blade/vane tips and the shroud can be decreased as low as possible while preventing and/or decreasing friction between the blade/vane tips and the shroud 958. In one example implementation, the clearance between the surface of the shroud 958 facing the secondary impeller blade/vanes 954 during operation of compressor may be less than 50 micrometers (μm). In another example implementation, the clearance between the surface of the shroud 958 facing the secondary impeller blade/vanes 954 during operation of the compressor may be between 5 μm and 15 μm. In yet another implementation, the aforementioned clearance may be between 7 μm and 13 μm or more preferably approximately 10 μm. In general, it is desirable to decrease the aforementioned clearance as much as possible while reducing or eliminating friction due to contact between the blade/vanes 954 and the shroud 958.

In one aspect of the disclosure, the blades and/or vanes 954 of the secondary impeller 915 may be configured to provide a pressure field and/or aerodynamic thrust during rotation of the secondary impeller 915. The pressure field and/or aerodynamic thrust may counteract the biasing force FF causing the desired clearance between the blade/vane 954 tips and the shroud 958 during operation of the compressor 910. The pressure field loading capacity of the blades/vanes of the secondary impeller 915 and working fluid can be calculated by solving Navier-Stokes Equations for thin fluid film as described above with respect to FIG. 4. The aforementioned calculation can be simplified to the example Reynolds Equations 1 and Equation 2 described above with respect to FIG. 4.

Using Navier-Stokes Equations or example Reynolds Equations provided above, the geometry of the secondary impeller 915 and/or the blades/vanes 954 of the secondary impeller 915 can be optimized to achieve a desired clearance (e.g., the clearances described above) between the blades/vanes 954 and the shroud 958 while reducing or eliminating friction due to contact between the blades/vanes 954 and the shroud 958. Additional details of a blade/vane profile that may be relevant to the described secondary impeller 915 are described with respect to FIG. 5 above.

The compressor 410 of FIG. 9 may have a primary impeller 913 and a secondary impeller 915 operatively connected to a drivetrain (e.g., drivetrain 217 in FIG. 2). The primary impeller 913 may have a plurality of vanes or blades 934 configured to compress fluid supplied at a primary impeller inlet 912. It is noted that, in FIG. 9, only a subset of blades and/or vanes 934 are provided with reference numbers to prevent obstruction of other features in the figure. Further, it is noted that the blade and/or vane structure is not limited to the structure shown in FIG. 9 and may include any number of blades and/or vanes in any configuration without departing from the scope of the disclosure. The primary impeller inlet 912 may for example have one or more electronically or hydraulically controlled vanes (hidden from view in FIG. 9, see, e.g., reference 214 in FIG. 2) for controlling the flow of fluid into the primary impeller 913.

The vanes and/or blades 934 of the rotating primary impeller 913 are configured to compress the fluid sucked into the primary impeller inlet 912 and are configured discharge the compressed fluid at the primary impeller tip 932 into the outlet path 920. In one example, the primary impeller 913 may have a dynamic shroud 933 that is fixed to and configured to rotate with the primary impeller 913. In one example, the dynamic shroud may be a unitary structure and formed and/or machined as a single component. In another example, the dynamic shroud 933 may be fastened to or otherwise connected to the blades and/or vanes 934 of the dynamic shroud 933. The primary impeller 913 may be sealed to prevent fluid leakage via one or more seals 936 configured to sealingly engage with a sealing surface, surfaces, or sealing features 937 of the primary impeller 913. In one aspect, the one or more seals may for example comprise one or more radial or circular labyrinth seal(s) or any type of seal that provides a tortuous path to prevent leakage of fluid at the rotational interface of the primary impeller. In some examples, one or more of the seal(s) 936 and/or features 937 may prevent or resist passage of fluid while providing non-contact with an opposing surface or sealing feature(s) by controlling the passage of fluid through a variety of chambers within the seal by centrifugal motion and/or the formation of controlled fluid vortices.

The secondary impeller 915 may be operatively connected to the primary impeller 913. In some examples, the primary impeller 913 and the secondary impeller 915 may be joined or formed/machined as a unitary structure. In other aspects, the secondary impeller 915 may be connectable to or otherwise connected with the primary impeller 913 via one or more mechanical fasteners or joining methods. For example, the secondary impeller 915 may be mounted to or otherwise joined to the dynamic shroud 933 of the primary impeller 913.

The secondary impeller 915 may have a significantly reduced size, diameter, and/or volume compared to the primary impeller 913. The secondary impeller 915 may have a significantly reduced size, diameter, and/or volume compared to the primary impeller 913. For example, the outer or major diameter of the secondary impeller 915 may be between 70% and 85% of the primary impeller 913 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 915 may be between 40% and 95% of the main or primary impeller 913 outer or major diameter. In another example, the outer diameter or major diameter of the secondary impeller 913 may be between 40% and 85% of the main impeller 913 outer or major diameter.

In addition the blades of the secondary impeller may have a blade profile with a higher degree of twist compared to the primary impeller. For example, the blades of the secondary impeller 915 may have a blade outlet angle 5 degrees (5°) to 30° lower than the primary impeller 913. For reference, a blade outlet angle of 90° angle would include blades that are purely radial at the outlet. The reference example 90° outlet angle would provide the highest lift capacity but tend to reduce flow capacity range of the impeller.

FIG. 10A is a partial view of an impeller 1015a with impeller blades 1034a with a 45° outlet angle. An Impeller with blades at 45° at the outlet may be usable with the secondary impeller(s) described herein and may provide a good tradeoff. In one aspect, an impeller with blades at 45° at the exit may be used as the primary impeller(s) described herein. Lower blade exit angles (i.e., lower than) 45° are generally not used in compressors due to a decreased lift capacity. However, in the case of a secondary impeller(s) describe herein, a lower blade exit angle (e.g., lower than) 45° may be preferred in one aspect and may provide the required lift for the economizer cycle described herein.

FIG. 10B shows one example of an impeller with blades 1034b at an angle of 30° at the outlet. The 30° blade angle at the outlet is usable with aspects of the secondary impeller(s) described herein. As can be seen when comparing the 45° exit blade profile of FIG. 10A with the 30° exit blade profile of FIG. 10b, as the blade exit angle decreases the blades look more “spiraled” or “twisted.” Thus, in one example, the blade profile of the secondary impeller described herein may be more spiraled or twisted than the blade profile of the primary impeller.

As noted above, the aforementioned blade angle profiles are provided as non-limiting examples. As noted above, the blade outlet profile of the secondary impeller(s) described herein may be 5°-30° lower than the blade outlet profile of the primary impeller(s) described herein.

FIGS. 10C and 10D show a partial view of a first example impeller inlet with blades 1014c having a first blade inlet profile and a second example impeller inlet 1015d with blades 1034d having a second blade inlet profile. In some examples the first blade inlet profile or secondary blade inlet profile are usable with either one of or any combination of the blade exit profiles shown in FIGS. 10A and 10b.

In addition, the secondary impeller blades and/or vanes may be adjusted accordingly to reduce the compression ratio of the economizer fluid flow through the compressor compared to the suction line fluid flow through the compressor. One notable difference between the impeller of FIG. 4 and FIG. 9 is that the outer diameter or major diameter of the secondary impeller 915 is larger than the secondary impeller 413 of FIG. 4. In other words, the difference in diameters between primary impeller 913 and secondary impeller 915 is smaller than the difference in diameter between the primary impeller 413 and secondary impeller 415. In some examples, the primary impeller 913 and secondary impeller 915 may have the same diameter or major diameter.

In one example, the number of veins, profile of the veins and/or height of the veins of the secondary impeller 915 may be modified to ensure that the proper pressure ratio or compression ratio is achieved. As noted above, the pressure ratio may be optimized using equations 1, 2, and 3 noted above. For example, the height of the impeller blades 954 of the secondary impeller 915 may be less that the height of the impeller blades 454 of the secondary impeller 415 in FIG. 4.

Utilizing a primary impeller 913 and a secondary impeller 915 as described throughout this disclosure allows for optimization of the geometries of the primary impeller 913 and/or the secondary impeller 915 to allow for an overall increased efficiency of a refrigeration cycle while reducing the impact of or any decrease in efficiency that the compression of the economizer fluid flow by the secondary impeller 915 has on the compression of the suction line fluid flow by the primary impeller 913. For example, since the vapor from the economizer or intercooler (e.g., economizer fluid flow from flash gas line 112 in FIG. 1) requires less kinetic energy than is needed to properly compress the gas received from the evaporator via the suction line (e.g., suction line 107 in FIG. 1), the secondary impeller may be adjusted (e.g., in any one or any combination of diameter, size, volume, and/or blade/vane configuration) to reduce the impact on efficiency that compression of the economizer fluid flow has on the compression of the suction line fluid flow. In one example, a tip diameter of the secondary impeller 915 may be chosen to adjust the compression ratio of suction line fluid flow and/or the economizer fluid and the economizer pressure. In some examples, the distance from RR or major diameter from one primary impeller tip 932 to an impeller tip on the opposite side of the primary impeller 913 may be greater than a distance from axis RR or major diameter from one secondary impeller tip 952 to an impeller tip on the opposite side of the secondary impeller 915.

As shown in FIG. 9, the secondary impeller 915 may have one or more blades and/or vanes 954. It is noted that, in FIG. 9, only a subset of blades and/or vanes 954 are provided with reference numbers to prevent obstruction of other features in the figure. Further, while a specific vane/blade configuration is shown in the figure, it is noted that any vane and/or blade configuration may be utilized without departing from the scope of the disclosure. The vanes and/or blades 954 of the rotating secondary impeller 915 are configured to compress the fluid received from the economizer and/or intercooler (for example via flash gas line 112 in FIG. 1), which may be interchangeably referred to herein as economizer fluid flow. In one example, the economizer fluid flow may be provided to an inner cavity 918 of the compressor casing.

As denoted by the example dashed lines with arrowheads representing flows in FIG. 9, the economizer fluid may be sucked into or otherwise be provided to a secondary impeller inlet 935, and the compressed economizer fluid flow fluid may exit the secondary impeller 915 at the secondary impeller outlet tip or tips 952. In one example, the secondary impeller outlet tip or tips 952 may be in fluid communication with the outlet path 920 thus causing the compressed economizer flow and suction line fluid flow to mix or partially mix before exiting the compressor 910. In one example, the compressor 910 may include a separation wall 971 at the outlet of the secondary impeller 915. The separation wall 971 may be utilized to delay the mixing of the economizer fluid flow and the suction line fluid flow until the suction line fluid flow velocity has decreased. Delaying the mixing of the economizer fluid flow and the suction line fluid flow until the velocity difference between the suction line fluid flow and the economizer fluid flow has decreased reduces any mixing losses that may otherwise occur. Any one or any combination of the separation wall length, geometry, and/or the positional relationship of the separation wall with respect to the primary impeller tip(s) 932 and/or the secondary impeller tip(s) 952 may be optimized to decrease the velocity difference between the economizer fluid flow and the suction line fluid flow at the divergence of the compressed economizer fluid flow and the suction line fluid flow at the outlet path 920.

Additional aspects of the disclosure are described in the clauses that follow:

• • Clause 1. A compressor system, comprising: a rotatable impeller configured to increase a pressure of one or more fluids from a compressor inlet to a compressor outlet, wherein the rotatable impeller further comprises: a primary impeller section having a first set of impeller blades configured to pressurize a first fluid source; and a secondary impeller section with a second set of impeller blades configured to pressurize a second fluid source.
  • Clause 2. The compressor system of clause 1, further comprising: a primary impeller inlet configured to receive the first fluid source; a secondary impeller inlet configured to receive the second fluid source; and an outlet path in which the pressurized first fluid and second fluid are configured to mix before exiting the compressor.
  • Clause 3. The compressor system of any of the preceding clauses, wherein a first major diameter of the first set of impeller blades is larger than a second major diameter of the second set of impeller blades.
  • Clause 4. The compressor system of any of the preceding clauses, further comprising a primary shroud connected to both the first set of impeller blades and the second set of impeller blades.
  • Clause 5. The compressor system of any of the preceding clauses, further comprising a secondary shroud connected to the second set of impeller blades.
  • Clause 6. The compressor system of any of the preceding clauses, further comprising a floating shroud configured to maintain a clearance with the second set of impeller blades when the second set of impeller blades rotate with respect to the floating shroud.
  • Clause 7. The compressor system of any of the preceding clauses, wherein the floating shroud has a biasing member configured to provide a biasing force towards the second set of impeller blades.
  • Clause 8. The compressor system of any of the preceding clauses, wherein the second set of impeller blades are configured to provide a pressure field when the rotatable impeller is rotated, wherein the pressure field counteracts the biasing force of the biasing member.
  • Clause 9. The compressor system of any of the preceding clauses, wherein the pressure field counteracts the biasing force until equilibrium between the biasing force and the pressure field is achieved, wherein at the equilibrium the clearance between the second set of impeller blades and the floating shroud is less than or equal to 50 micrometers (μm).
  • Clause 10. The compressor system of any of the preceding clauses, further comprising a single impeller hub shared by the primary impeller and the secondary impeller.
  • Clause 11. A refrigeration system, comprising: a condenser; an evaporator; an expansion apparatus; an economizer or intercooler; and a compressor, wherein the compressor comprises: a rotatable impeller with a primary impeller section having a first set of impeller blades configured to pressurize a refrigerant received from the evaporator, and a secondary impeller section with a second set of impeller blades configured to pressurize refrigerant received from the economizer or intercooler.
  • Clause 12. The system of clause 11, wherein the primary impeller inlet is configured to receive refrigerant from the evaporator and the secondary impeller inlet is configured to receive refrigerant from the economizer or intercooler, wherein after compression, the compressor is configured to mix the refrigerant from the evaporator and the refrigerant from the economizer or intercooler before the refrigerant is supplied to the condenser.
  • Clause 13. The system of any of the preceding clauses, wherein a first major diameter of the first set of impeller blades is larger than a second major diameter of the second set of impeller blades.
  • Clause 14. The system of any of the preceding clauses, further comprising a primary shroud connected to both wherein the first set of impeller blades and the second set of impeller blades.
  • Clause 15. The system of any of the preceding clauses, further comprising a secondary shroud connected to the second set of impeller blades.
  • Clause 16. The system of any of the preceding clauses, wherein the compressor further comprises a floating shroud configured to maintain a clearance with the second set of impeller blades when the second set of impeller blades rotate with respect to the shroud.
  • Clause 17. The system of any of the preceding clauses, wherein the floating shroud has a biasing member configured to provide a biasing force towards the second set of impeller blades.
  • Clause 18. The system of any of the preceding clauses, wherein the second set of impeller blades are configured to provide a pressure field when the rotatable impeller is rotated, wherein the pressure field counteracts the biasing force of the biasing member.
  • Clause 19. The system of any of the preceding clauses, wherein the pressure field counteracts the biasing force until equilibrium between the biasing force and the pressure field is achieved, wherein at the equilibrium the clearance between the second set of impeller blades and the floating shroud is less than or equal to 50 micrometers (μm).
  • Clause 20. An impeller usable with a centrifugal compressor, comprising: an impeller hub with a shrouded primary impeller section having a first set of impeller blades and configured to pressurize a first fluid source; and a secondary impeller section with a second set of impeller blades connected to the shrouded primary impeller section and configured to pressurize a second fluid source.
  • Clause 21. The impeller of clause 20, further comprising a secondary shroud connected to the second set of impeller blades.

This written description uses examples to disclose aspects of the invention, including the preferred embodiments, and also to enable any person skilled in the art to practice the aspects thereof, including making and using any devices or systems and performing any incorporated methods. The patentable scope of these aspects is defined by the claims, and may include other examples that occur to those skilled in the art. Such other examples are intended to be within the scope of the claims if they have structural elements that do not differ from the literal language of the claims, or if they include equivalent structural elements with insubstantial differences from the literal language of the claims. Aspects from the various embodiments described, as well as other known equivalents for each such aspect, can be mixed and matched by one of ordinary skill in the art to construct additional embodiments and techniques in accordance with principles of this application.