Refrigeration system with high speed rotary pressure exchanger

Abstract

To increase efficiency of a refrigeration system.SOLUTION: A refrigeration system includes a rotary pressure exchanger fluidly coupled to a low pressure branch and a high pressure branch. The rotary pressure exchanger is configured to receive a refrigerant at high pressure from the high pressure branch, to receive a refrigerant at low pressure from the low pressure branch, and to exchange pressure between the refrigerant at high pressure and the refrigerant at low pressure. A first exiting stream from the rotary pressure exchanger includes the refrigerant at high pressure in the supercritical state or the subcritical state and a second exiting stream from the rotary pressure exchanger includes the refrigerant at low pressure in the liquid state or the two-phase mixture of liquid and vapor.SELECTED DRAWING: Figure 27

Description

[Technical Field]

[0001] This section is intended to introduce the reader to various aspects of art that may be related to various aspects of the present invention, which are described and / or claimed below. This discussion is believed to be helpful in providing the reader with background information to facilitate a better understanding of the various aspects of the present invention. As such, it should be understood that these statements are to be read in this light, and not as admissions of prior art. [Background technology]

[0002] Due to enforcement by government environmental agencies, much of the world is now being forced to transition to zero-global-warming refrigeration systems such as supercritical carbon dioxide refrigeration. While supercritical carbon dioxide systems work well in relatively cool climates, such as most of Europe and North America, they face drawbacks in hotter climates because the coefficient of performance (a measure of efficiency) of these systems decreases as the ambient temperature of the surrounding environment increases, resulting in higher electrical costs per unit of cooling delivered. This is because supercritical carbon dioxide systems require much higher pressures to operate (approximately 10,342 kPa (1500 psi) or more) than HFC / CFC-based systems (approximately 1,379 to 2,068.4 kPa (200 to 300 psi)). To raise the refrigerant above its critical pressure, very high differential pressure compressors are used. The large pressure ratio across the compressor consumes more electrical energy. This problem is magnified in hotter climates because the refrigerant temperature at the inlet of the chiller must be raised high enough to allow heat rejection to the hotter ambient environment. This is done by increasing the pressure ratio across the compressor ever higher, thus requiring more electricity for the compressor, which in turn increases the cost of electricity per unit of cooling performed. Increasing the efficiency of refrigeration systems (e.g., supercritical carbon dioxide refrigeration systems) can reduce the operating costs of refrigeration systems and increase their availability, while helping to reduce global warming. Summary of the Invention

[0003] Certain embodiments consistent in scope with the disclosed subject matter are summarized below. These embodiments are not intended to limit the scope of the disclosure, but rather merely to provide a brief summary of some embodiments of the disclosed subject matter. Indeed, the disclosure may encompass a variety of forms that may be similar to or different from the embodiments set forth below.

[0004] In one embodiment, a refrigeration system is provided. The refrigeration system includes a high-pressure branch for circulating a refrigerant at a high pressure therethrough. The refrigeration system also includes a gas cooler or condenser disposed along the high-pressure branch, where the high-pressure branch is configured to reject heat from the refrigerant at a high pressure to the ambient environment through the gas cooler or condenser, the high-pressure refrigerant being in a supercritical or subcritical state. The refrigeration system further includes a low-pressure branch for circulating a refrigerant at a low pressure therethrough. The refrigeration system also includes an evaporator disposed along the low-pressure branch, where the low-pressure branch is configured to absorb heat from the ambient environment through the evaporator into a low-pressure refrigerant in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. The refrigeration system also includes a compressor or pump configured to increase the pressure of the refrigerant from the low pressure to the high pressure. The refrigeration system further includes a rotary pressure exchanger fluidly coupled to the low-pressure branch and the high-pressure branch, wherein the rotary pressure exchanger is configured to receive a high-pressure refrigerant from the high-pressure branch and a low-pressure refrigerant from the low-pressure branch and to exchange pressure between the high-pressure refrigerant and the low-pressure refrigerant, wherein a first exit stream from the rotary pressure exchanger includes the high-pressure refrigerant in a supercritical or subcritical state and a second exit stream from the rotary pressure exchanger includes the low-pressure refrigerant in a liquid state or a two-phase mixture of liquid and vapor.

[0005] In one embodiment, a refrigeration system is provided. The refrigeration system includes a high-pressure branch for circulating a refrigerant at a high pressure therethrough. The refrigeration system includes a gas cooler or condenser disposed along the high-pressure branch, where the high-pressure branch is configured to reject heat from the refrigerant at high pressure to the ambient environment through the gas cooler or condenser, the high-pressure refrigerant being in a supercritical or subcritical state. The refrigeration system also includes a low-pressure branch for circulating a refrigerant at a low pressure therethrough. The refrigeration system further includes a first evaporator disposed along the low-pressure branch, where the first evaporator is configured to operate at a first temperature and the low-pressure branch is configured to absorb heat from the ambient environment through the evaporator into a low-pressure refrigerant, the low-pressure refrigerant being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. The refrigeration system further includes a first intermediate-pressure branch for circulating a refrigerant at a first intermediate pressure therethrough. The refrigeration system also includes a second evaporator disposed along the first intermediate pressure branch, wherein the second evaporator is configured to operate at a second temperature higher than the first temperature. The refrigeration system further includes a second intermediate pressure branch for circulating a refrigerant therethrough at a second intermediate pressure, wherein the first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the respective pressures of the refrigerant in the low-pressure branch and the second intermediate pressure branch, the first intermediate pressure of the refrigerant in the first intermediate pressure branch is equal to a saturation pressure in the second evaporator, and the second intermediate pressure of the refrigerant in the second intermediate pressure branch is between the respective pressures of the refrigerant in the high-pressure branch and the first intermediate pressure branch.The refrigeration system further includes a flash tank operating at a second intermediate pressure and configured to separate the liquid-vapor two-phase mixture of refrigerant into a pure liquid and a pure vapor; and a rotary pressure exchanger fluidly coupled to the second intermediate pressure branch and the high pressure branch, the rotary pressure exchanger configured to receive a high pressure refrigerant from the high pressure branch and a second intermediate pressure refrigerant in a vapor state, a liquid state, or a two-phase mixture of liquid and vapor from the second intermediate pressure branch and to exchange pressures between the high pressure refrigerant and the second intermediate pressure refrigerant, wherein a first exit stream from the rotary pressure exchanger includes the high pressure refrigerant in a supercritical or subcritical state, and a second exit stream from the rotary pressure exchanger includes the second intermediate pressure refrigerant in a liquid state or a two-phase mixture of liquid and vapor.

[0006] In one embodiment, a refrigeration system is provided. The refrigeration system includes a high-pressure branch for circulating a refrigerant at a high pressure therethrough. The refrigeration system also includes a gas cooler or condenser disposed along the high-pressure branch, where the high-pressure branch is configured to reject heat from the refrigerant at a high pressure to the ambient environment through the gas cooler or condenser, the high-pressure refrigerant being in a supercritical or subcritical state. The refrigeration system further includes a second low-pressure branch for circulating a refrigerant at a low pressure therethrough. The refrigeration system also includes a first evaporator disposed along the low-pressure branch, where the first evaporator is configured to operate at a first temperature, and where the low-pressure branch is configured to absorb heat from the ambient environment through the evaporator into a low-pressure refrigerant, the low-pressure refrigerant being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. The refrigeration system also includes an intermediate-pressure branch for circulating a refrigerant at an intermediate pressure therethrough. The refrigeration system also includes a second evaporator disposed along the intermediate-pressure branch, where the second evaporator is configured to operate at a second temperature higher than the first temperature. An intermediate pressure of the refrigerant in the intermediate-pressure branch is between the respective pressures of the refrigerant in the high-pressure branch and the low-pressure branch, and the intermediate pressure of the refrigerant in the intermediate-pressure branch is equal to a saturation pressure in the second evaporator. The refrigeration system further includes a flash tank operating at the intermediate pressure and configured to separate the two-phase liquid-vapor mixture of refrigerant into a pure liquid and a pure vapor. The refrigeration system further includes a rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high pressure branch, wherein the rotary pressure exchanger is configured to receive a high pressure refrigerant from the high pressure branch and an intermediate pressure refrigerant in a vapor state, a liquid state, or a two-phase mixture of liquid and vapor from the intermediate pressure branch and to exchange pressure between the high pressure refrigerant and the intermediate pressure refrigerant, wherein a first exit stream from the rotary pressure exchanger includes the high pressure refrigerant in a supercritical state or a subcritical state, and a second exit stream from the rotary pressure exchanger includes the intermediate pressure refrigerant in a liquid state or a two-phase mixture of liquid and vapor.

[0007] The various features, aspects and advantages of the present invention will become better understood from the following detailed description when taken in conjunction with the accompanying drawings, in which like characters represent like parts throughout. [Brief explanation of the drawings]

[0008] [Figure 1] FIG. 1 is a phase diagram of carbon dioxide.

[0009] [Figure 2] FIG. 2 is a schematic diagram of one embodiment of a refrigeration system having a rotary pressure exchanger or rotary liquid piston compressor (LPC).

[0010] [Figure 3] FIG. 3 is a temperature-entropy diagram illustrating the thermodynamic processes in a refrigeration system utilizing a Joule-Thomson expansion valve versus the refrigeration system of FIG.

[0011] [Figure 4] FIG. 4 is a pressure-enthalpy diagram of the thermodynamic processes in a refrigeration system utilizing a Joule-Thomson expansion valve versus the refrigeration system of FIG.

[0012] [Figure 5] FIG. 5 is an exploded perspective view of one embodiment of a rotary pressure exchanger or rotary LPC.

[0013] [Figure 6] FIG. 6 is an exploded perspective view of one embodiment of a rotary pressure exchanger or rotary LPC in a first operating position.

[0014] [Figure 7] FIG. 7 is an exploded perspective view of one embodiment of a rotary pressure exchanger or rotary LPC in a second operating position.

[0015] [Figure 8]FIG. 8 is an exploded perspective view of one embodiment of a rotary pressure exchanger or rotary LPC in a third operational position.

[0016] [Figure 9] FIG. 9 is an exploded perspective view of one embodiment of a rotary pressure exchanger or rotary LPC in a fourth operational position.

[0017] [Figure 10] FIG. 10 is an exploded view of one embodiment of a rotor with a barrier system.

[0018] [Figure 11] FIG. 11 is a cross-sectional view of one embodiment of a rotor with a barrier system.

[0019] [Figure 12] FIG. 12 is a cross-sectional view of one embodiment of a rotor with a barrier system.

[0020] [Figure 13] FIG. 13 is a cross-sectional view of one embodiment of a rotor with a barrier system.

[0021] [Figure 14] FIG. 14 is a cross-sectional view of one embodiment of the barrier taken along line 14-14 of FIG.

[0022] [Figure 15] FIG. 15 is a cross-sectional view of one embodiment of the barrier taken along line 14-14 of FIG.

[0023] [Figure 16] FIG. 16 is a cross-sectional view of one embodiment of a rotary pressure exchanger or rotary liquid piston compressor with a cooling system.

[0024] [Figure 17] FIG. 17 is a cross-sectional view of one embodiment of a rotary pressure exchanger or rotary liquid piston compressor with a heating system.

[0025] [Figure 18] FIG. 18 is a schematic diagram of one embodiment of a refrigeration system in a supermarket refrigeration system architecture.

[0026] [Figure 19] FIG. 19 is a schematic diagram of one embodiment of a refrigeration system in an alternative supermarket refrigeration system architecture.

[0027] [Figure 20] FIG. 20 is a schematic diagram of one embodiment of a control system for controlling the movement of motive and working fluids within an RLPC.

[0028] [Figure 21] FIG. 21 is a schematic diagram of one embodiment of a control system for controlling the movement of motive and working fluids within an RLPC.

[0029] [Figure 22A] FIG. 22A is a schematic diagram of one embodiment of a refrigeration system with a rotary pressure exchanger or rotary liquid piston compressor (LPC) (e.g., one with a low flow, high differential pressure (DP) leak pump and a low DP, high flow circulating pump instead of a bulk flow compressor).

[0030] [Figure 22B] FIG. 22B is a schematic diagram of an embodiment of a refrigeration system having a rotary pressure exchanger or rotary liquid piston compressor (LPC) (eg, one with a leak compressor instead of a bulk flow compressor).

[0031] [Figure 23] FIG. 23 is a temperature-entropy diagram of the thermodynamic processes within the refrigeration system of FIG.

[0032] [Figure 24]FIG. 24 is a pressure-enthalpy diagram of the thermodynamic processes within the refrigeration system of FIG.

[0033] [Figure 25] FIG. 25 is a schematic diagram of one embodiment of a refrigeration system having a rotary pressure exchanger or rotary liquid piston compressor (LPC), such as one with a leakage compressor instead of a bulk flow compressor and an additional low DP recycle compressor (e.g., a blower).

[0034] [Figure 26] FIG. 26 is a schematic diagram of one embodiment of a refrigeration system in a supermarket refrigeration system architecture (eg, with an expansion valve).

[0035] [Figure 27] FIG. 27 is a schematic diagram of one embodiment of a refrigeration system in an alternative supermarket refrigeration system architecture (eg, with an expansion valve). DETAILED DESCRIPTION OF THE INVENTION

[0036] One or more specific embodiments of the present invention are described below. These described embodiments are merely examples of the present invention. Moreover, in an effort to provide a concise description of these exemplary embodiments, all features of an actual implementation may not be described herein. It should be recognized that, as with any engineering or design project, many implementation-specific decisions must be made in developing any such actual implementation to achieve the developer's specific end goals, such as compliance with system-related and business-related constraints, which may vary from implementation to implementation. Moreover, it should be recognized that such a development effort may be complex and time-consuming, but would be considered a routine undertaking of design, fabrication, and manufacture for those of ordinary skill in the art having the benefit of this disclosure. Furthermore, it should be recognized that features of different embodiments disclosed herein can be combined with each other unless otherwise indicated.

[0037] The following discussion describes a refrigeration system (e.g., a supercritical carbon dioxide refrigeration system) that utilizes a rotary pressure exchanger or rotary liquid piston compressor or rotary liquid piston pump instead of a Joule-Thomson expansion valve. As described below, the refrigeration system can operate more efficiently by increasing the cooling capacity of the refrigeration system while recovering a large portion of the pressure energy that would otherwise be lost if a Joule-Thomson expansion valve were used. Replacing a Joule-Thomson expansion valve with a rotary pressure exchanger increases efficiency due to the elimination of both energy destruction and entropy generation that occurs within the expansion valve, which results in total losses of up to 40 percent in typical refrigeration systems. Furthermore, replacing a Joule-Thomson expansion valve with a rotary pressure exchanger increases efficiency by changing the expansion process from an isenthalpic (i.e., constant enthalpy) process across the expansion valve to an isentropic or near-isentropic (i.e., constant entropy) process across the rotary pressure exchanger. In some embodiments, a rotary pressure exchanger can also replace the function of a bulk flow compressor. This allows one or more low-DP recycle compressors (blowers) or recycle pumps to be used in place of bulk-flow high-DP compressors and maintain flow rates within the refrigeration system (e.g., to overcome small pressure losses). These low-DP recycle compressors may consume significantly less energy (e.g., 10 times less) than bulk-flow compressors. Replacing both the Joule-Thomson expansion valve and the bulk-flow compressor with a rotary pressure exchanger eliminates two of the largest sources of inefficiency within the refrigeration system while reducing power consumption and electricity costs. Furthermore, using a rotary pressure exchanger in place of the expansion valve and / or bulk-flow compressor may increase the usability of the refrigeration system in other environments (e.g., warmer environments).At warmer ambient temperatures (e.g., 50°C), the compressor pressure ratio is altered (significantly increasing the pressure required at the compressor outlet), significantly reducing the cycle efficiency (i.e., coefficient of performance) by as much as 60 percent compared to the optimum temperature (e.g., 35°C). The rotary pressure exchanger mitigates the adverse effects of warmer ambient temperatures on the required compressor work, the refrigeration system cooling capacity, and the refrigeration system's coefficient of performance.

[0038] During operation, a rotary pressure exchanger or rotary liquid piston compressor or pump may or may not perfectly equalize the pressure between the first and second fluids. Thus, a rotary liquid piston compressor or pump may operate isobarically or substantially isobarically (e.g., where the pressures of the first and second fluids are equalized within approximately + / - 1, 2, 3, 4, 5, 6, 7, 8, 9, or 10 percent of each other). A rotary liquid piston compressor or pump may generally be defined as a device that transfers fluid pressure between a high-pressure inlet stream and a low-pressure inlet stream with an efficiency of greater than approximately 50%, 60%, 70%, 80%, or 90%.

[0039] Figure 1 shows a phase diagram for carbon dioxide. A phase diagram represents the equilibrium limits of various phases within a chemical system as a function of temperature and pressure. Phase diagram 2 in Figure 1 illustrates how carbon dioxide changes phase (e.g., gas (vapor), liquid, solid, supercritical) as temperature and pressure change. In addition to illustrating when carbon dioxide can exist as a gas or vapor, liquid, and solid, phase diagram 2 also illustrates when carbon dioxide can transform into a supercritical fluid. A compound becomes a supercritical fluid when subjected to pressures and temperatures above its critical point. The critical point is the point at which the surface tension (meniscus) that separates the liquid and gas phases of a substance disappears, making the two phases indistinguishable. Within the supercritical region, the fluid exhibits special properties. These properties may include a liquid-like (e.g., an order of magnitude higher) density, specific heat, viscosity, and speed of sound through the gas.

[0040] FIG. 2 is a schematic diagram of one embodiment of a refrigeration system 800 using a fluid in a supercritical state (e.g., a supercritical carbon dioxide refrigeration system). While the refrigeration system 800 is described as using carbon dioxide, other refrigerants are possible. The use of a rotary pressure exchanger or rotary liquid compressor 802 (labeled PX in the figure) as described below in place of an expansion valve (e.g., a Joule-Thomson expansion valve) in the refrigeration system 800 allows the refrigeration system 800 to operate more efficiently by increasing the cooling capacity of the refrigeration system 800 while recovering much of the pressure energy that would be lost if a Joule-Thomson expansion valve were used. In some embodiments, the rotary pressure exchanger can replace the function of a bulk flow compressor, thus allowing one or more (significantly more energy-efficient) low-DP recycle compressors or pumps to be used in place of the bulk flow compressor. For example, supercritical carbon dioxide refrigeration systems must operate at much greater pressures (approximately 10,342 kPa (1500 psi) or greater), which creates a large pressure ratio across the compressor (a very high differential pressure compressor) and results in more electrical energy being consumed. Replacing the expansion valve with a rotary pressure exchanger allows nearly all of the pressure drop to be recovered in the rotary pressure exchanger and utilized to compress the flow coming from the evaporator rather than sending the flow to the main compressor. Thus, the electrical demand on the compressor can be significantly reduced or eliminated. Refrigeration system 800 using a rotary pressure exchanger instead of a Joule-Thomson expansion valve and / or bulk flow compressor can be used in a variety of applications, including supermarket refrigeration systems, heating, ventilation, and / or air conditioning (HVAC) systems, refrigeration for natural liquefied gas systems, industrial refrigeration for the chemical processing industry, battery technology (e.g., creating thermal energy storage systems for solar or wind power generation using a combined refrigeration and power generation cycle), aquariums, polar habitat research systems, and any other system where refrigeration is used.

[0041] As depicted, refrigeration system 800 includes a first fluid loop (e.g., a high-pressure branch) 804 for circulating a high-pressure refrigerant (e.g., carbon dioxide) and a second fluid loop (e.g., a low-pressure branch) 806 for circulating a low-pressure refrigerant (e.g., carbon dioxide) at a lower pressure than the high-pressure branch 804. The first fluid loop 804 includes a heat exchanger 808 (e.g., a gas cooler / condenser) and a rotary pressure exchanger 802. The heat exchanger 808 rejects heat from the high-pressure refrigerant to the ambient environment. While the gas cooler is described below for use with supercritical high-pressure refrigerants (e.g., carbon dioxide), in some embodiments, the condenser can be used with subcritical high-pressure refrigerants (e.g., carbon dioxide). The subcritical state for a refrigerant is below the critical point (specifically, between the critical point and the triple point). The second fluid loop 806 includes a heat exchanger 810 (e.g., a cooling or heat load such as an evaporator) and the rotary pressure exchanger 802. The heat exchanger 810 absorbs heat from the ambient environment into a low-pressure refrigerant. The low-pressure refrigerant in the low-pressure branch 806 may be in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. Both fluid loops 804, 806 are fluidly coupled to a compressor 812 (e.g., a bulk flow compressor). The compressor 812 converts (by increasing the temperature and pressure) the superheated gaseous carbon dioxide received from the evaporator 810 to carbon dioxide in a supercritical state, which is supplied to the gas cooler 808. In some embodiments, as described in further detail below, the compressor 812 may be replaced by one or more low-DP circulating compressors or pumps to overcome small pressure losses and maintain fluid flow within the system 800. Generally, along the first fluid loop 804, the gas cooler 808 receives the carbon dioxide in a supercritical state, which it then supplies to the rotary pressure exchanger 802 after some cooling (e.g., at the high-pressure inlet 822). Along second fluid loop 804, evaporator 810 supplies a first portion of the superheated gaseous carbon dioxide to a low pressure inlet 813 of rotary pressure exchanger 802 and supplies a second portion of the superheated gaseous carbon dioxide to compressor 812. Rotary pressure exchanger 802 exchanges pressure between the carbon dioxide in a supercritical state and the superheated gaseous carbon dioxide.The carbon dioxide in a supercritical state is converted into a two-phase liquid / vapor mixture within rotary pressure exchanger 802 and exits through low pressure outlet 824, which is supplied to evaporator 810. Rotary pressure exchanger 802 also increases the pressure and temperature of the superheated gaseous carbon dioxide, converting it to carbon dioxide in a supercritical state, which exits rotary pressure exchanger 802 through high pressure outlet 815, where it is supplied to gas cooler 808. As illustrated in Figure 2, the carbon dioxide in a supercritical state exiting rotary pressure exchanger 802 may be combined with carbon dioxide supplied to gas cooler 808 from compressor 812.

[0042] The thermodynamic processes occurring within refrigeration system 800 (e.g., in the context of a refrigeration system using a Joule-Thomson expansion valve) are described in more detail with reference to FIGS. 3 and 4. FIGS. 3 and 4 illustrate a temperature-entropy (TS) diagram 814 and a pressure-entropy (PH) diagram 816, respectively, to show the thermodynamic processes occurring in the four major components of refrigeration system 800 compared to a refrigeration system including a Joule-Thomson expansion valve. Point 1 represents the compressor inlet 818 (see FIG. 2). Point 2 represents the compressor outlet 819 and the gas cooler inlet 820. Point 3 represents the gas cooler outlet 830 and the expansion valve inlet (in a refrigeration system with a Joule-Thomson expansion valve) or the high-pressure inlet 822 of the rotary liquid compressor 802. Point 4 represents the expansion valve outlet or the low-pressure outlet 824 of the rotary liquid compressor 802 (labeled PX in FIGS. 3 and 4) and the evaporator inlet 826. As illustrated in FIGS. 3 and 4 , the compressor 812 increases the pressure, thus raising the temperature of the refrigerant working fluid (e.g., carbon dioxide), above the ambient temperature, where it can reject heat to the external, warmer ambient. This occurs within the gas cooler 808. Unlike a conventional condenser where the temperature remains constant throughout most of the heat exchange process inside the two-phase dome on the TS diagram within the gas cooler 808 of a supercritical carbon dioxide system because the carbon dioxide is in a supercritical state, there is no phase boundary, and the carbon dioxide is above the two-phase dome 828. Therefore, the temperature drops as the carbon dioxide rejects heat to the ambient. The higher the ambient temperature, the greater the pressure ratio across the compressor 812, and the greater the system pressure. The carbon dioxide leaving the gas cooler outlet 830 at point 3 now travels through the expansion valve (in a refrigeration system with a Joule-Thomson expansion valve) and follows the constant enthalpy process (3 → 4h) within the valve, as shown by curve 832. On the pH diagram 816, curve 832 is a straight vertical line (because it is an isenthalpic process). As a result, the carbon dioxide enters the two-phase dome 828 and becomes an equilibrium mixture of liquid and gas.The exact mass fraction of the liquid is determined by the point where 4h (i.e., curve 832) intersects with the constant pressure horizontal line 834, which represents the evaporator pressure. The two-phase mixture now continues through evaporator 810, where the liquid carbon dioxide absorbs more heat and becomes a saturated vapor at the outlet 836 of evaporator 810. The fluid entering compressor 812 is therefore in pure vapor (gas) phase.

[0043] Now consider a system with a rotary pressure exchanger 802 replacing a Joule-Thomson expansion valve as shown in FIG. 2. As illustrated in FIGS. 3 and 4, carbon dioxide in a supercritical state at a gas cooler outlet 830 enters the rotary pressure exchanger 802 at a high-pressure inlet port 822, undergoes isentropic or near-isentropic expansion (e.g., 85 percent isentropic efficiency), and exits at a low-pressure outlet port 824 of the rotary pressure exchanger 802 as two-phase gas-liquid carbon dioxide. This process is illustrated by curve 835 on the TS and PH diagrams 814, 816. As illustrated, curve 835 (obtained using the rotary pressure exchanger 802) lies to the left of curve 832 (obtained using an expansion valve), indicating that the amount or percentage of liquid content in the two-phase liquid is higher than in the case of an expansion valve (point 4 on the PH diagram 816). h This means that the expansion through the rotary pressure exchanger 802 (point 4 on the PH diagram 816) is greater than that through the evaporator 810 (point 1 on the PH diagram 816). Due to the higher liquid content, the absorption capacity of the refrigerant (e.g., carbon dioxide) is greater in the evaporator 810. Therefore, for the same pressure and temperature boundary conditions set by the environmental conditions, the cooling capacity of the refrigeration system 800 increases when the rotary pressure exchanger 802 is used instead of the Joule-Thomson expansion valve. The location of point 4 on the PH diagram 816 represents a complete isentropic expansion process (e.g., a 100 percent isentropic expansion coefficient). The two-phase carbon dioxide at point 4 will then absorb heat in the evaporator 810 (process 4 → 1). The length 838 (4 h-4) is the additional cooling capacity provided by the system 800 using the rotary pressure exchanger 802 compared to a typical one using a Joule-Thomson expansion valve (segment length 834, which is the difference between the enthalpy at point 1 and the enthalpy at point 4h). This is one of the primary advantages provided by incorporating the rotary pressure exchanger 802 within the refrigeration cycle.

[0044] Another advantage provided by using rotary pressure exchanger 802 within the refrigeration cycle is that carbon dioxide enters rotary pressure exchanger 802 from evaporator 80 as superheated gaseous carbon dioxide (at low pressure inlet 813) and undergoes isentropic or near isentropic (e.g., 85 percent isentropic efficiency) compression as shown by dashed line 842 (i.e., process 1 → 2 s ) second fluid stream. This process is similar to the isentropic process 1 → 2 occurring inside compressor 812. Because nearly all of the compression occurs inside rotary pressure exchanger 802, in some embodiments it is possible to completely or partially eliminate main compressor 812. For example, in this case compressor 812 may be replaced by a very low differential pressure gas blower or circulation pump that consumes very little work (because the enthalpy change across it is very small). This creates a significant advantage in the efficiency of the refrigeration cycle, as can be seen from the following equation for coefficient of performance (COP) (i.e., the standard measure of the efficiency of a refrigeration cycle):

number

[0045] FIG. 5 is an exploded perspective view of an embodiment of a rotary pressure exchanger or rotary liquid piston compressor 40 (rotary LPC) (e.g., rotary pressure exchanger 802 in FIG. 2 ) capable of transferring pressure and / or work between a first fluid (e.g., supercritical carbon dioxide circulating in a first fluid loop 804) and a second fluid (e.g., superheated gaseous carbon dioxide circulating in a second fluid loop 806) with minimal mixing of the fluids. Rotary LPC 40 may include a generally cylindrical body portion 42 including a sleeve 44 (e.g., a rotor sleeve) and a rotor 46. Rotary LPC 40 may also include two end caps 48 and 50 including manifolds 52 and 54, respectively. Manifold 52 includes respective inlet and outlet ports 56 and 58, while manifold 54 includes respective inlet and outlet ports 60 and 62. During operation, the inlet ports 56, 60 allow first and second fluids to enter the rotary LPC 40 and exchange pressures, while the outlet ports 58, 62 allow the first and second fluids to subsequently exit the rotary LPC 40. During operation, the inlet port 56 can receive a first fluid at high pressure, and after exchanging pressures, the outlet port 58 can be used to direct the first fluid at low pressure out of the rotary LPC 40. Similarly, the inlet port 60 can receive a second fluid at low pressure, and the outlet port 62 can be used to direct the second fluid at high pressure out of the rotary LPC 40. The end caps 48 and 50 include respective end covers 64 and 66 disposed within the respective manifolds 52 and 54 that enable fluid-tight contact with the rotor 46. The rotor 46 can be cylindrical and disposed within the sleeve 44, allowing the rotor 46 to rotate about an axis 68. The rotor 46 may have a plurality of channels 70 extending substantially longitudinally therethrough, with openings 72 and 74 at each end arranged symmetrically about the longitudinal axis 68. The openings 72 and 74 in the rotor 46 are arranged to be in fluid communication with inlet and outlet apertures 76 and 78; and 80 and 82 in the end covers 64 and 66 such that the channels 70 are exposed to high and low pressure fluids during rotation.As illustrated, the entrance and exit apertures 76 and 78; 80 and 82 may be designed in the form of an arc or a segment of a circle (eg, C-shaped).

[0046] In some embodiments, a controller using sensor feedback (e.g., revolutions per minute measured through a tachometer or optical encoder or volumetric flow rate measured through a flow meter) can control the degree of mixing between the first and second fluids within the rotary LPC 40, which can be used to improve the operability of the fluid handling system. For example, by varying the volumetric flow rates of the first and second fluids entering the rotary LPC 40, a plant operator (e.g., a system operator) can control the amount of fluid mixing within the rotary liquid piston compressor 10. Furthermore, by varying the rotational speed of the rotor 46, the operator can similarly control mixing. Three characteristics of the rotary LPC 40 that affect mixing are (1) the aspect ratio of the rotor channel 70, (2) the exposure duration between the first and second fluids, and (3) the creation of a fluidic barrier (e.g., interface) between the first and second fluids within the rotor channel 70. First, the rotor channel 70 is generally long and narrow, which stabilizes the flow rate within the rotating LPC 40. Furthermore, the first and second fluids may move through the channel 70 in a plug flow regime with minimal axial mixing. Second, in some embodiments, the speed of the rotor 46 reduces contact between the first and second fluids. For example, the speed of the rotor 46 may reduce the contact time between the first and second fluids to approximately 0.15 seconds, 0.10 seconds, or less than 0.05 seconds. Third, only a small portion of the rotor channel 70 is used for pressure exchange between the first and second fluids. Therefore, a fixed volume of fluid remains within the channel 70 as a barrier between the first and second fluids. All of these mechanisms may limit mixing within the rotating LPC 40. Moreover, in some embodiments, the rotary LPC 40 may be designed to operate with a full or partial internal piston or other barrier that isolates the first and second fluids while allowing pressure transfer.

[0047] 6-9 are exploded views of one embodiment of the rotary LPC 40 illustrating the sequence of positions of a single rotor channel 70 within the rotor 46 as the channel 70 rotates through one complete cycle. FIGS. 6-9 are simplified views of the rotary LPC 40 showing one rotor channel 70, with the channel 70 shown as having a circular cross-sectional shape. In other embodiments, the rotary LPC 40 can include multiple channels 70 having the same or different cross-sectional shapes (e.g., circular, square, square, polygonal, etc.). Accordingly, FIGS. 6-9 are simplified for illustrative purposes, and other embodiments of the rotary LPC 40 can have configurations different from those shown in FIGS. 6-9. As described in more detail below, the rotary LPC 40 facilitates pressure exchange between the first and second fluids by allowing the first and second fluids to briefly contact each other within the rotor 46. In some embodiments, this exchange occurs at a rate that results in limited mixing of the first and second fluids. More specifically, the speed of the pressure wave traveling through the rotor channel 70 (as soon as the channel is exposed to the aperture 76), the rate of diffusion of the fluid, and the rotational speed of the rotor 46 determine whether and to what extent any mixing occurs.

[0048] In FIG. 6 , the channel opening 72 is in a first position. In the first position, the channel opening 72 is in fluid communication with an aperture 78 in the end cover 64 and thus the manifold 52, while the opposing channel opening 74 is in fluid communication with an aperture 82 in the end cover 66 and thus the manifold 54. As discussed below, the rotor 46 may rotate in a clockwise direction, indicated by arrow 84. During operation, a low-pressure second fluid 86 passes through the end cover 66 and enters the channel 70, where it contacts the first fluid 88 at a dynamic fluid interface 90. The second fluid 86 then moves out of the channel 70, through the end cover 64, and out of the rotating LPC 40, displacing the first fluid 88. However, due to the short contact time, there is minimal mixing between the second fluid 86 and the first fluid 88.

[0049] 7, channel 70 has rotated clockwise through an approximately 90-degree arc. In this position, opening 74 (e.g., outlet) is no longer in fluid communication with apertures 80 and 82 in end cover 66, and opening 72 is no longer in fluid communication with apertures 76 and 78 in end cover 64. Thus, low-pressure second fluid 86 is temporarily stored within channel 70.

[0050] In Figure 8, channel 70 has rotated through approximately a 60 degree arc from the position shown in Figure 7. Opening 74 is now in fluid communication with aperture 80 in end cover 66, and opening 72 of channel 70 is now in fluid communication with aperture 76 in end cover 64. In this position, a high pressure first fluid 88 enters and pressurizes a low pressure second fluid 86, forcing second fluid 86 out of rotor channel 70 through aperture 80.

[0051] In Figure 9, channel 70 has rotated through an arc of approximately 270 degrees from the position shown in Figure 6. In this position, opening 74 is no longer in fluid communication with apertures 80 and 82 in end cover 66, and opening 72 is no longer in fluid communication with apertures 76 and 78 in end cover 64. Therefore, first fluid 88 is no longer pressurized and is temporarily stored within channel 70 until rotor 46 rotates another 90 degrees to begin the cycle again.

[0052] FIG. 10 is an exploded view of one embodiment of the rotor 46 having a barrier system 100. As described above, rotation of the rotor 46 allows for pressure transfer between the first and second fluids. The rotary liquid piston compressor 10 includes the barrier system 100 to prevent mixing between the first fluid / motive fluid and the second fluid / supercritical fluid within the power generation system 4. As illustrated, the rotor 46 includes a first rotor section 102 and a second rotor section 104 coupled together. By including the rotor 46 with the first and second rotor sections 102, 104, the rotor 46 can receive and retain the barrier system 100 within the rotor 46. As illustrated, the first rotor section 102 includes an end face 106 having an aperture 108 that receives a bolt 110. Bolts 110 pass through these apertures 108 and into apertures 112 in the second rotor section 104, joining the first and second sections 102, 104 of the rotor 46. The barrier system 100 is installed between these rotor sections 102, 104, securely fastening the barrier system 100 to the rotor 46.

[0053] The barrier system 100 includes a plate 114 having a plurality of barriers 116 coupled to it. The barriers 116 are foldable diaphragms that prevent contact / mixing between the first and second fluids as they exchange pressure within the channels 70 of the rotor 46. As discussed below, the barriers 116 expand and contract as pressure is transferred between the first and second fluids. To couple the plate 114 to the rotor 46, the plate 114 may include a plurality of apertures 118 that align with the apertures 108 in the first rotor section 102 and the apertures 112 in the second rotor section 104. The apertures 118 receive the bolts 110 and reduce or prevent lateral movement of the plate 114 when the first rotor section 102 is coupled to the second rotor section 104. In some embodiments, the apertures 108 on the first rotor section 102, the apertures 112 on the second rotor section 104, and the apertures 118 on the plate 114 can be located on one or more diameters (e.g., inner and outer diameters). In this manner, the first rotor section 102 and the second rotor section 104 can evenly compress the plate 114 when coupled together. In some embodiments, the barriers 116 may not be coupled to or supported by the plate 114. Rather, each barrier 116 may be individually coupled to the rotor 46.

[0054] As illustrated, the first rotor section 102 defines a length 120, and the second rotor section 104 defines a length 122. By varying the lengths 120 and 122, the rotor 46 allows for the barrier system 100 to be installed at different locations within the channel 70 along the length of the rotor 46. In this manner, the rotary liquid piston compressor 10 can be adapted to meet various operating conditions. For example, differences in the densities and mass flow rates of the two fluids and the rotational speed of the rotor 46, among other factors, can affect how far the first and second fluids can enter the channel 70 of the rotor 46 to exchange pressure. Thus, varying the lengths 120 and 122 of the first and second rotor sections 102 and 104 of the rotor 46 allows for the barrier system 100 to be installed at a location (e.g., midway through the rotor 46) that promotes pressure exchange between the first and second fluids.

[0055] In some embodiments, refrigeration system 800 can modify the fluids circulating in first and second loops 804 and 806 to prevent mixing within rotary liquid piston compressor 802. For example, refrigeration system 800 can use an ionic fluid in first loop 804 that can prevent diffusion and solubility of the supercritical fluid in another fluid that is in a different phase, or in other words, prevent mixing with the supercritical fluid. Modification of the fluids in refrigeration system 800 can also be used in combination with barrier system 100, which provides redundant resistance to mixing of the fluids within rotary liquid piston compressor 802.

[0056] FIG. 11 is a cross-sectional view of one embodiment of the rotor 46 having a barrier system 100. As described above, the barrier system 100 can include plates 114 and barriers 116. These barriers 116 are positioned within the channels 70 to block mixing / contact between the first and second fluids while still allowing pressure transfer. The barriers 116 expand and contract to facilitate pressure transfer. As illustrated in FIG. 11 , a first barrier 140 of the barriers 116 is in an expanded position. During operation, the first barrier 140 expands as the first fluid 142 flows into the rotor 46 and into the first barrier 140. As the first barrier 140 expands, it pressurizes the second fluid 144 and forces it out of the rotor 46. Simultaneously, the second barrier 146 can be in a contracted state as the second fluid 144 enters the rotor 46 in preparation for pressurization. The barrier 116 includes a plurality of corrugations 148 (e.g., 1, 2, 3, 4, 5 or more) joined together by ribs 150. It is these elastic corrugations 148 that allow the barrier 116 to expand in volume as the pressurized first fluid 142 flows into the rotor 46. As discussed below, the barrier 116 may be made of one or more materials that provide the tensile strength, elongation percentage, and chemical resistance to function with supercritical fluids (e.g., carbon dioxide).

[0057] FIG. 12 is a cross-sectional view of one embodiment of a rotor 46 having a barrier system 100. As illustrated in FIG. 12, a first barrier 140 of the plurality of barriers 116 is in an expanded position. During operation, the first barrier 140 expands as a first fluid 142 flows into the rotor 46 and into the first barrier 140. As the first barrier 140 expands, it contracts, pressurizing the second fluid 144 and forcing it out of the rotor 46. To reduce stress within the barrier 116, the barrier system 100 may include a spring 160. The spring 160 may be coupled to an end 162 (e.g., an end portion, an end face) of the barrier 116 and to the plate 114. During operation, the spring 160 expands as pressure within the barrier 116 increases, causing the barrier 116 to expand in an axial direction 164. Because the spring 160 absorbs forces as the barrier 116 expands, the spring 160 can prevent or reduce over-expansion of the barrier 116. The spring 160 can also increase the life of the barrier 116 as the barrier 116 repeatedly expands and contracts during operation of the power generation system 4. The spring can also provide a more controlled rate of expansion of the barrier 116.

[0058] In some embodiments, the spring 160 can be coupled to an exterior surface 168 of the barrier 116 and / or located on the outside of the barrier 116. In other embodiments, the spring 160 can be coupled to an interior surface 170 and / or located on the interior of the barrier 116 (i.e., within the membrane of the barrier 116). In still other embodiments, the barrier system 100 can include a spring 160 on both the exterior and interior of the barrier 116. The spring 160 can also be coupled to the rotor 46 instead of being coupled to the plate 114. For example, the spring 160 can be supported by sandwiching a portion of the spring 160 between the first rotor section 102 and the second rotor section 104 of the rotor 46.

[0059] FIG. 13 is a cross-sectional view of one embodiment of a rotor 46 having a barrier system 100. In FIG. 13, the barrier system 100 includes a planar barrier 190. As illustrated, the planar barrier 190 extends across the channel 70 (e.g., generally transverse to the longitudinal axis of the channel 70) rather than axially into the channel 70 as with the barrier 116 described above. During operation, the planar barrier 190 prevents mixing / contact between the first and second fluids 142, 144 while still allowing pressure transfer. To facilitate pressure transfer, the planar barrier 190 expands and contracts under pressure. As illustrated in FIG. 13, a first planar barrier 192 of the plurality of planar barriers 190 is in an expanded position. The first planar barrier 192 expands as the first fluid 142 flows into the rotor 46 and into the first planar barrier 192. As the first planar barrier 192 expands under the pressure of the first fluid 142, it contacts the second fluid 144, pressurizing it and forcing it out of the rotor 46. The second planar barrier 194 may likewise simultaneously contract as the second fluid 144 enters the rotor 46 in preparation for being pressurized. The barrier 116 includes a plurality of interlocking corrugations 196 (e.g., 1, 2, 3, 4, 5 or more). It is these elastic corrugations 148 that expand as the pressurized first fluid 142 flows into the rotor 46 and contract when the pressure is released.

[0060] FIG. 14 is a cross-sectional view of one embodiment of the barrier 116 along line 14-14 in FIG. 11. Barrier 116 and barrier 190 may be made of one or more materials that provide the tensile strength, elongation percentage, and chemical resistance to function with supercritical fluids (e.g., carbon dioxide). For example, barriers 116, 190 may include a high-stretch ratio elastomeric material, such as ethylene propylene, silicone, nitrile, or neoprene. The high-stretch ratio capabilities of these materials enable barriers 116, 119 to absorb pressure from first fluid 142 and transfer it to second fluid 144. In some embodiments, barriers 116, 119 may include multiple layers (e.g., 1, 2, 3, 4, 5, or more layers) of high-stretch ratio material sandwiched between layers of high-strength fabric to combine high-stretch ratio and high-strength properties. For example, barriers 116, 119 may include two elastomeric layers 210 overlapping fabric layer 212. In operation, the elastomer layer 210 can provide chemical resistance as well as high elongation rate capability, while the fabric layer 212 can increase the overall tensile strength of the barrier 116 , 190 .

[0061] FIG. 15 is a cross-sectional view of one embodiment of the barrier 116, 190 taken along line 14-14 in FIG. 11. As discussed above, the barrier 116, 190 may be made of one or more materials that provide the tensile strength, elongation percentage, and chemical resistance to function in supercritical fluids (e.g., the temperatures and pressures of the supercritical fluid). In some embodiments, the barrier 116, 119 may include multiple layers (e.g., 1, 2, 3, 4, 5, or more layers) to combine the properties of different materials. For example, the barrier 116, 119 may include two elastomer layers 210 (e.g., ethylene propylene, silicone, nitrile, neoprene, etc.) overlapping a fabric layer 212. During operation, the elastomer layer 210 may provide chemical resistance and high elongation ratio capabilities, while the fabric layer 212 increases the tensile strength of the barrier 116, 190. Additionally, one or more of the layers 210 may include a coating 214. Coating 214 can be a chemically resistant coating that resists reaction with the first fluid and / or the second fluid. For example, layer 210 can include a coating 214 on its outermost surface 216 that chemically protects layer 210 from the supercritical fluid.

[0062] FIG. 16 is a cross-sectional view of one embodiment of a rotary liquid piston compressor 10 (e.g., a rotary LPC) with a refrigeration system 240 (i.e., a thermal management system). In some embodiments, the refrigeration system 240 can include a heat exchanger fabricated with microchannels surrounding the rotary liquid piston compressor. As explained above in the description of FIG. 1, fluids change phase as temperature and pressure change. At pressures and temperatures above their critical points, the fluid becomes a supercritical fluid. The refrigeration system 800 uses the fluid (e.g., carbon dioxide) in its supercritical state / phase for refrigeration purposes due to the unique properties of supercritical fluids (e.g., liquid-like density and gas-like viscosity). By controlling the temperature within the rotary liquid piston compressor 10 with the refrigeration system 240, the refrigeration system 240 can prevent a phase change from a supercritical fluid to a gas phase within the rotary liquid piston compressor 802. Additionally, the cooling system 240 may also facilitate the removal of energy as heat is generated during compression of the supercritical fluid, allowing for substantially isothermal compression, which is a thermodynamically more efficient mode of compression. As discussed above, the cooling system 240 may include microchannels that provide a high surface area per unit volume to facilitate a heat transfer coefficient between the walls of the rotary liquid piston compressor 802 and the cooling fluid circulating through the cooling system 240.

[0063] The cooling system 240 includes a cooling jacket 242 that surrounds at least a portion of the rotary liquid piston compressor housing 244. The cooling jacket 242 may include a plurality of conduits 246 that encase the housing 244. These conduits 246 may be microconduits having diameters between 0.05 mm and 0.5 mm. By including the microconduits, the cooling system 240 can increase the cooling surface area for controlling the temperature of the supercritical fluid within the rotary liquid piston compressor 10. The conduits 246 may be arranged in multiple rows (e.g., 1, 2, 3, 4, 5, or more) and / or multiple columns (e.g., 1, 2, 3, 4, 5, or more). Each conduit 246 may be fluidly coupled to every other conduit 246, or the cooling system 240 may be coupled to a subset of the conduits 246. For example, all of the conduits 246 in a row may be coupled to other conduits 246 in that row but not to conduits 246 in other rows. In some embodiments, each conduit 246 may be fluidly coupled to other conduits 246 in the same column but not to conduits 246 in different columns. In some embodiments, the conduits 246 may be surrounded by a housing or covering 247. The housing or covering 247 may be made of a material that blocks and resists heat transfer, such as polystyrene, fiberglass wool, or various types of foam. The flow rate of the cooling fluid through the conduits 246 may be controlled by a controller 248. The controller 248 may include a processor 250 and a memory 252. For example, the processor 250 may be a microprocessor that executes software to control the operation of the actuators 98. Processor 250 may include multiple microprocessors, one or more "general purpose" microprocessors, one or more special purpose microprocessors, and / or one or more application specific integrated circuits (ASICS), or some combination thereof. For example, processor 250 may include one or more reduced instruction set (RISC) processors.

[0064] The memory 252 may include volatile memory, such as random access memory (RAM), and / or non-volatile memory, such as read-only memory (ROM). The memory 252 may store a variety of information and may be used for a variety of purposes. For example, the memory 252 may store processor-executable instructions, such as firmware or software, for execution by the processor 250. The memory may include ROM, flash memory, a hard drive, or any other optical, magnetic, or solid-state storage medium, or a combination thereof. The memory may store data, instructions, and any other suitable data.

[0065] During operation, the controller 248 receives feedback from one or more sensors 254 (e.g., temperature sensors, pressure sensors) that directly or indirectly detect the temperature and / or pressure of the supercritical fluid. Using the feedback from the sensors 254, the controller 248 controls the flow rate of the cooling fluid from a cooling fluid source 256 (e.g., a chiller system, an air conditioning system).

[0066] 17 is a cross-sectional view of one embodiment of a rotary liquid piston compressor 802 (RLPC) having a heating system 280 (i.e., thermal management system). During operation, the heating system 280 can control the temperature of the fluid (i.e., supercritical fluid) circulating through the rotary liquid piston compressor 802. By controlling the temperature, the heating system 280 can prevent or reduce condensation of the fluid and / or dry ice formation due to non-isentropic expansion.

[0067] The heating system 280 includes a heating jacket 282 that surrounds at least a portion of the rotary liquid piston compressor housing 244. The heating jacket 282 may include a plurality of conduits or cables 284 that encase the housing 244. These conduits or cables 284 enable temperature control of the supercritical fluid. For example, the conduits 284 may carry a heating fluid that transfers heat to the supercritical fluid. In some embodiments, the cable(s) 284 (e.g., coils) carry an electrical current that generates heat due to the electrical resistance of the cable 284. The conduits 246 may in turn be surrounded by a housing or covering 286. The housing or covering 286 may be made of a material that blocks and resists heat transfer, such as polystyrene, fiberglass wool, or various types of foam.

[0068] The flow rate of the heating fluid or current through the conduit or cable 284 is controlled by the controller 248. During operation, the controller 248 can receive feedback from one or more sensors 254 (e.g., temperature sensors, pressure sensors) that directly or indirectly detect the temperature and / or pressure of the supercritical fluid. For example, the sensor 254 may be located in direct contact with the supercritical fluid (e.g., inside a cavity that stores the supercritical fluid). In some embodiments, the sensor 254 may be located within the housing 244, sleeve 44, or end cover 64, 66. As the material around the sensor 254 responds to changes in the temperature and / or pressure of the supercritical fluid, the sensor 254 detects the change and communicates the change to the controller 248, which then correlates this to the actual temperature and / or pressure of the supercritical fluid. Using the feedback from the sensor 254, the controller 248 controls the flow rate of the heating fluid from a heating fluid source 288 (e.g., a boiler) through the conduit 284. Similarly, if the heating system 280 is an electrical resistance heating system, the controller 248 may control the flow rate of current through the cable 284 in response to feedback from one or more of the sensors 254 .

[0069] Figures 18 and 19 show two example supermarket system architectures 300, 302 that use a rotary pressure exchanger-based supercritical carbon dioxide refrigeration system rather than traditional Joule-Thomson expansion valve-based refrigeration. In the first architecture 300 (Figure 18), the two-phase low-pressure outlet (e.g., a carbon dioxide gas / liquid mixture) from the rotary pressure exchanger 304 (via low-pressure outlet 305) passes through a flash tank 306, which separates the gas and liquid phases. The carbon dioxide liquid phase is transported to low-temperature (e.g., approximately -20 degrees Celsius (C)) and medium-temperature (e.g., approximately -4°C) heat loads / evaporators 308, 310 (e.g., the freezer and refrigerator sections of a supermarket, respectively), where it absorbs heat and becomes superheated. Because it is a pure liquid phase rather than a two-phase gas / liquid phase, it has greater heat absorption (i.e., cooling) capacity. Flow control valves 312, 314 (e.g., in response to control signals from a controller) can adjust the flow rate of liquid carbon dioxide to the respective heat loads 308, 310. The superheated carbon dioxide vapor from the freezer section 308 then travels to a low-temperature compressor 316, where it is recombined, at the same pressure, with the superheated carbon dioxide vapor from the refrigerator section 310 and the separated superheated gas-phase carbon dioxide that was separated from the gas / liquid mixture in the flash tank 306. A control valve 318 (e.g., a flash gas control valve) can adjust the flow rate of the superheated gaseous carbon dioxide flowing from the flash tank 306 (e.g., in response to control signals from a controller). This recombined superheated gaseous carbon dioxide then enters the rotary pressure exchanger 304 at a low-pressure inlet port 320, where it is compressed to the highest pressure in the system (e.g., approximately 10,342 kPa (1500 psi) or approximately 14,479 kPa (2100 psi), depending on system requirements), and converted to supercritical carbon dioxide. The supercritical carbon dioxide exits the rotary pressure exchanger 304 (via high pressure outlet 322) and proceeds to heat exchanger 324 at the highest pressure, where it rejects heat to the environment and cools. In some embodiments, heat exchanger 324 is a gas condenser used with subcritical carbon dioxide.From the gas cooler 324, the supercritical carbon dioxide flows to the high pressure inlet 326 of the rotary pressure exchanger 304. The small pressure rise required to overcome the flow resistance in the system and the small pressure differential in the rotary pressure exchanger 304 can be provided by using a small compressor 328 (e.g., a low DP recycle compressor) (shown between the path from the rotary pressure exchanger 304 and the gas cooler 324) with significantly less energy consumption than a conventional compressor.

[0070] A heat exchanger 324 is disposed along the high-pressure branch for circulating carbon dioxide in a supercritical or subcritical state at high pressure. A low-temperature evaporator 308 and a low-temperature compressor 316 are disposed along the low-pressure branch for circulating carbon dioxide at low pressure (i.e., a pressure lower than that in the high-pressure branch) in a liquid, gaseous, or vapor state, or a two-phase mixture of liquid and vapor. An intermediate-temperature evaporator 310 and a valve 314 are disposed along the intermediate-pressure branch for circulating refrigerant at an intermediate pressure between the respective pressures of the refrigerants in the high-pressure and low-pressure branches. The intermediate pressure of the refrigerant in the intermediate-pressure branch is equal to the saturation pressure in the evaporator 310. The refrigerant exiting the flash tank 306 and flowing directly to the inlet 320 of the rotary pressure exchanger 304 is at an intermediate pressure. The rotary pressure exchanger 304 is therefore fluidly coupled to the intermediate-pressure and high-pressure branches. The rotary pressure exchanger 304 receives high-pressure refrigerant from the high-pressure branch and intermediate-pressure refrigerant in the vapor state, liquid state, or two-phase mixture of liquid and vapor from the intermediate-pressure branch, and exchanges pressure between the high-pressure refrigerant and the intermediate-pressure refrigerant. A first exit stream of refrigerant at high pressure in a supercritical or subcritical state and a second exit stream of refrigerant at intermediate pressure in a liquid state or two-phase mixture of liquid and vapor exits the rotary pressure exchanger.

[0071] In the second architecture 302 (FIG. 19), only the separated gaseous carbon dioxide from the flash tank is sent again at low-pressure inlet 320 through rotary pressure exchanger 304 and compressed to the highest pressure in the system. Superheated gaseous carbon dioxide from freezer section 308 and refrigerator section 310, respectively, flows to low-temperature compressor 316 and medium-temperature compressor 330. The low-temperature compressor outlet is combined with the superheated gaseous carbon dioxide from refrigerator section 310 before going to medium-temperature compressor 330. The medium-temperature compressor outlet (e.g., supercritical carbon dioxide) combines with the supercritical carbon dioxide exiting rotary pressure exchanger 304 (via high-pressure outlet 322), where it is combined with the already compressed low-temperature compressor outlet and the medium-temperature compressor outlet (superheated gaseous carbon dioxide at the same pressure as flash tank 306) before going through gas cooler 324. Such an architecture may have advantages in some refrigeration scenarios.

[0072] A heat exchanger 324 is disposed along the high-pressure branch for circulating carbon dioxide in a supercritical or subcritical state at high pressure. A low-temperature evaporator 308 and a low-temperature compressor 316 are disposed along the low-pressure branch for circulating carbon dioxide at low pressure (i.e., a pressure lower than that in the high-pressure branch) in a liquid state, a gaseous or vapor state, or a two-phase mixture of liquid and vapor. An intermediate-temperature evaporator 310 and a valve 314 are disposed along the first intermediate-pressure branch for circulating a refrigerant at a first intermediate pressure between the respective pressures of the refrigerants in the low-pressure branch and the second intermediate-pressure branch. The second intermediate-pressure branch is disposed between the flash tank 306 and the rotary pressure exchanger 304. The first intermediate pressure of the refrigerant in the intermediate-pressure branch is equal to the saturation pressure in the evaporator 310. The refrigerant exiting the flash tank 306 and flowing directly to the inlet 320 of the rotary pressure exchanger 304 is at a second intermediate pressure between the respective pressures in the high-pressure branch and the first intermediate-pressure branch. Thus, rotary pressure exchanger 304 is fluidly coupled to the second intermediate-pressure branch and the high-pressure branch. Rotary pressure exchanger 304 receives high-pressure refrigerant from the high-pressure branch and receives second intermediate-pressure refrigerant in a vapor state, liquid state, or in the form of a two-phase mixture of liquid and vapor from the second intermediate-pressure branch, and exchanges pressure between the high-pressure refrigerant and the second intermediate-pressure refrigerant. Exiting the rotary pressure exchanger are a first exit stream of refrigerant at high pressure in a supercritical or subcritical state, and a second exit stream of refrigerant at a second intermediate pressure in a liquid state or in the form of a two-phase mixture of liquid and vapor.

[0073] FIG. 20 is a schematic diagram of one embodiment of a control system 570 controlling the movement of a fluid (e.g., supercritical carbon dioxide, superheated gaseous carbon dioxide) within a rotary pressure exchanger or rotary liquid piston compressor 572. As discussed above, a rotary liquid piston compressor can be used to exchange energy between two fluids. For example, the rotary liquid piston compressor 572 can be used to exchange energy between two fluids within the refrigeration system described above. In response to the flow rate of the working fluid 580, the control system 570 can control the flow rate of the superheated gaseous carbon dioxide 574 into the rotary liquid piston compressor 572 in order to reduce transfer of the superheated gaseous carbon dioxide 574 or two-phase gas / liquid carbon dioxide mixture 575 within the fluid loop 576 and / or prevent it from entering the fluid loop 578 that circulates the working fluid (i.e., the superheated carbon dioxide 580). That is, by controlling the flow rate of the superheated gaseous carbon dioxide 574, the control system 570 can prevent and / or restrict the superheated gaseous carbon dioxide 574 from flowing completely through the rotary liquid piston compressor 572 (i.e., completely out of the channel 70 seen in FIG. 5 ) and into the working fluid loop 578.

[0074] To control the flow rate of superheated gaseous carbon dioxide 574, control system 570 includes a valve 582 that controls the amount of superheated gaseous carbon dioxide 574 entering rotary liquid piston compressor 572. Sensors 586 and 588 sense the respective flow rates of superheated gaseous carbon dioxide 574 and working fluid 580 and generate signals indicative of the flow rates. That is, sensors 586 and 588 measure the respective flow rates of superheated gaseous carbon dioxide 574 and working fluid 580 into rotary liquid piston compressor 572. Controller 584 receives and processes signals from sensors 586, 588 to detect the flow rates of superheated gaseous carbon dioxide 574 and working fluid 580.

[0075] In response to the detected flow rate, controller 584 controls valve 582 to prevent and / or reduce the transfer of superheated gaseous carbon dioxide 574 into working fluid loop 578. For example, if controller 584 detects a reduced flow rate via sensor 588, controller 584 can relate the reduced flow rate to how far the working fluid has penetrated in direction 590 into rotary liquid piston compressor 572. Thus, controller 584 can determine an associated flow rate of superheated gaseous carbon dioxide 574 into rotary liquid piston compressor 572 that will exit working fluid 580 in direction 592 from rotary liquid piston compressor 572 without exiting superheated gaseous carbon dioxide 574 from rotary liquid piston compressor 572 in direction 592. In other words, the controller 584 controls the valve 582 to prevent the flow of superheated gaseous carbon dioxide 574 into the working fluid loop 578 such that the flow rate of the working fluid 580 into the rotary liquid piston compressor 572 is greater than the flow rate of the superheated gaseous carbon dioxide 574.

[0076] As illustrated, the controller 584 may include a processor 594 and a memory 596. For example, the processor 594 may be a microprocessor that executes software to process signals from the sensors 586, 588 and control the operation of the valve 582 in response.

[0077] FIG. 21 is a schematic diagram of one embodiment of a control system 620 controlling the movement of a fluid (e.g., supercritical carbon dioxide, superheated gaseous carbon dioxide) within a rotary liquid piston compressor 622. As discussed above, a rotary liquid piston compressor or pump can be used to exchange energy between two fluids. For example, the rotary liquid piston compressor 622 can be used to exchange energy between two fluids within the refrigeration system described above. In response to the flow rate of the working fluid 630 and the flow rate of the superheated gaseous carbon dioxide 624, the control system 620 can control the axial travel distance of the carbon dioxide within the rotor channel within the rotary liquid piston compressor 622 in order to reduce the transfer of the superheated gaseous carbon dioxide 624 or two-phase gas / liquid carbon dioxide mixture 625 within the fluid loop 626 and / or prevent it from entering the working fluid loop 628, which circulates the working fluid 630 (e.g., superheated carbon dioxide). The control system 620 controls the movement of the driving fluid by slowing or accelerating the rotational speed of the rotor of the rotary liquid piston compressor 622. That is, by controlling the rotational speed, the control system 620 can prevent and / or limit the flow of superheated gaseous carbon dioxide 624 from flowing entirely through the rotary liquid piston compressor 622 (i.e., entirely through channel 70 seen in FIG. 5 ) and into the working fluid loop 628.

[0078] To reduce mixing of the working fluid 630 and the superheated gaseous carbon dioxide 624, the control system 620 includes a motor 632. The motor 632 controls the rotational speed of the rotor (e.g., rotor 46 shown in FIG. 5 ), and therefore the axial length of the rotor channels that the superheated gaseous carbon dioxide 624 can flow into. The faster the rotor rotates, the shorter the time that the superheated gaseous carbon dioxide and working fluid must flow into the rotor channels, and thus the shorter the axial length of the rotor channels occupied by the superheated gaseous carbon dioxide / process fluid. Similarly, the slower the rotor rotates, the longer the time that the superheated gaseous carbon dioxide and working fluid must flow into the rotor channels, and thus the longer the axial length of the rotor channels occupied by the superheated gaseous carbon dioxide / process fluid.

[0079] The control system 620 may include a variable frequency drive for controlling the motor and sensors 634 and 636 for sensing the respective flow rates of the superheated gaseous carbon dioxide 624 and the working fluid 630 and generating signals indicative of the flow rates. A controller 638 receives and processes the signals to detect the flow rates of the superheated gaseous carbon dioxide 624 and the working fluid 630. In response to the detected flow rates, the controller 638 sends commands to the variable frequency drive to control the speed of the motor 632 to prevent and / or reduce the transfer of the superheated gaseous carbon dioxide 624 into the working fluid loop 578. For example, if the controller 638 detects a decrease in the flow rate of the working fluid 630 using the sensor 636, the controller 638 can correlate this flow rate to how far the working fluid has traveled in the direction 640 into the channels of the rotary liquid piston compressor 622. Thus, the controller 638 can determine an associated speed of the motor 632 that forces the working fluid 630 out of the rotary liquid piston compressor 622 in the direction 642 without forcing the superheated gaseous carbon dioxide 624 out of the rotary liquid piston compressor 622 in the direction 642.

[0080] In response to a low instantaneous flow rate of the working fluid relative to the superheated gaseous carbon dioxide, the controller 638 controls the motor 632 through the variable frequency drive to increase the rotational speed (i.e., increase the revolutions per minute) of the rotary liquid piston compressor 622 to shorten the axial length that the superheated gaseous carbon dioxide 624 can travel within the channels of the rotary liquid piston compressor 622. Similarly, if the instantaneous flow rate of the working fluid 630 is too high compared to the driving fluid, the controller 638 reduces the rotational speed of the rotary liquid piston compressor 622 to increase the axial distance that the superheated gaseous carbon dioxide 624 can travel within the channels of the rotary liquid piston compressor 622 to force the working fluid 630 out of the rotary liquid piston compressor 622.

[0081] As illustrated, the controller 638 may include a processor 644 and a memory 646. For example, the processor 644 may be a microprocessor that executes software to process signals from the sensors 634, 636 and control the operation of the motor 632 in response.

[0082] As noted above, because nearly all of the compression occurs within the rotary pressure exchanger, in some embodiments, it is possible to completely or partially eliminate the main compressor (e.g., a bulk flow compressor). For example, the compressor can be replaced by a very low differential pressure gas blower or a circulation pump that consumes very little work (because the enthalpy change across it is very small). FIG. 22A is a schematic diagram of one embodiment of a refrigeration system 900 (e.g., a supercritical carbon dioxide refrigeration system) having a rotary pressure exchanger or rotary liquid piston compressor (LPC) 902 (e.g., with a low-flow, high-DP leakage pump and a low-DP, high-flow circulation pump instead of the bulk flow compressor). Generally, refrigeration system 900 is similar to refrigeration system 800 in FIG. 2.

[0083] As depicted, refrigeration system 900 includes a first fluid loop 904 and a second fluid loop 906. The first fluid loop (high-pressure loop) 904 includes a gas cooler or condenser 908, a high-pressure, high-flow, low-DP multi-phase circulation pump 909, and the high-pressure side of a rotary pressure exchanger 902. The second fluid loop (low-pressure loop) 906 includes an evaporator 910 (e.g., a cooling or heat load), a low-pressure, high-flow, low-DP multi-phase circulation pump 911, and the low-pressure side of the rotary pressure exchanger 902. The rotary pressure exchanger 902 fluidly couples the high-pressure and low-pressure loops 904, 906. Additionally, a multiphase leak pump 913, operating at a low flow rate but high DP, extracts any leakage from the pressure exchanger 902 at low pressure through a low-pressure outlet 920 and returns it to the high-pressure loop 904, just upstream of the high-pressure inlet 914 of the pressure exchanger 902. A multiphase pump 909 in the high-pressure loop 904 ensures that the required flow rate is maintained in the high-pressure loop 904 by overcoming the small pressure losses in the loop 904. Because there is not much pressure differential across the pump 909, the pump consumes very little energy. The flow entering this multiphase pump 909 is from the outlet 936 of the gas cooler / condenser 908 and could be in a supercritical state, a liquid state, or a two-phase mixture of liquid and vapor. Because there is not much pressure rise across the pump 909, the flow exiting the pump 909 is then considered to be at the same state as the inlet flow entering the high-pressure inlet 914 of the pressure exchanger 902. It is contemplated that the flow rate from the low pressure outlet 920 of the pressure transducer 902 may be in a two-phase liquid-vapor state or a pure liquid state.

[0084] A multiphase pump 911 in low-pressure loop 906 circulates this bulk, low-pressure flow of refrigerant through an evaporator 910 and to a low-pressure inlet 918 of pressure transducer 902. The multiphase pump 911 similarly has very little differential pressure across it (i.e., only enough to overcome any pressure losses in the system), and therefore, the pump 911 consumes very little energy compared to a conventional bulk-flow, high-pressure compressor. The low-pressure multiphase pump 911 circulates the flow through the evaporator 910, where it gains heat and converts itself to a pure vapor state or a higher vapor content two-phase liquid-vapor mixture. This high vapor content flow then enters the low-pressure inlet 918 of pressure transducer 902 and is pressurized to a high pressure. This, in turn, causes the temperature of the fluid to increase as well, according to the standard laws of thermodynamics. This high-pressure, higher-temperature fluid then exits the pressure transducer 902 through a high-pressure outlet 922. It is believed that the fluid exiting high-pressure outlet 922 may be in a supercritical state, or may exist in a subcritical state or as a liquid-vapor mixture with a high vapor content, depending on how the system is optimized. This high-pressure, high-temperature refrigerant then enters gas cooler / condenser 908 of high-pressure loop 904 and rejects heat to the ambient environment. By rejecting heat, the refrigerant either cools (if in a supercritical state) or changes phase to a liquid state. Multiphase pump 909 in high-pressure loop 904 then accepts this liquid refrigerant and circulates it through high-pressure loop 904 as previously described.

[0085] If there are no internal leaks within pressure transducer 902, high-pressure loop 904 will remain at a constant high pressure, and low-pressure loop 906 will remain at a constant low pressure. However, if there is an internal leak within pressure transducer 902 from the high-pressure side to the low-pressure side, there will be a net transfer of flow from high-pressure loop 904 to low-pressure loop 906. To account for this transfer and pump this leakage flow back into high-pressure loop 904, a third multiphase pump 913, a high-differential-pressure, low-flow leakage pump, is used. Pump 913 picks up any excess flow leaking into low-pressure loop 906 at low pressure and pumps it back to high-pressure loop 904, maintaining mass balance and pressure within each loop 904, 906. A three-way valve 915 is located within low-pressure loop 906 between the low-pressure outlet 920 of pressure transducer 902 and the inlet of low-pressure multiphase pump 911. Valve 915 splits the flow, allowing only the excess flow coming from the low-pressure outlet 920 of pressure transducer 902 to be directed to high-DP multiphase pump 913. Pump 913 also pumps any additional flow coming from low-pressure outlet 920 due to the compressibility of the refrigerant and the density differences between the four streams entering and leaving pressure transducer 902. Pump 913 also helps maintain a constant low pressure in low-pressure loop 906 and a constant high pressure in high-pressure loop 904. Another three-way valve 917 is positioned in high-pressure loop 904 between the outlet of high-pressure multiphase pump 909 and the high-pressure inlet 94 of pressure transducer 902. Valve 917 allows the leakage / excess flow from high-DP multiphase pump 913 to be combined with the high-pressure bulk flow coming from high-pressure multiphase pump 909 before being directed into high-pressure inlet 914 of pressure transducer 902. Although the pressure difference across the multiphase pump 913 is high, the flow rate it must pump is negligible (e.g., approximately 1-10 percent of the bulk flow rate going through either of the other two pumps 909, 911). Thus, the energy consumption of the pump 913 is relatively low as well.If the energy consumption of all three multiphase pumps 909, 911, 913 is added together, it would still be much lower than the energy consumption of a conventional compressor used to pressurize the entire bulk flow from the lowest pressure in the system (i.e., the evaporator pressure) to the highest pressure in the system (i.e., the condenser / gas cooler pressure), which is a major advantage of this configuration.

[0086] FIG. 22B demonstrates another embodiment of a bulk-flow compressor-less refrigeration system 923. It is similar to the system 900 shown in FIG. 22A, except that any excess flow exiting the low-pressure outlet 920 of the pressure transducer 902 (due to internal leaks in the pressure transducer 902 or due to compressibility and density differences between the four streams entering and exiting the pressure transducer 902 as previously described) is pumped through an evaporator 910 along with the valve low-pressure flow and converted to vapor before being compressed back into the high-pressure loop 904. Thus, the high-DP, low-flow, multiphase leakage pump 913 of FIG. 22A is replaced by a high-DP, low-flow, leakage compressor 925 as shown in FIG. 22B. The leakage compressor 925 compresses the excess flow from a low-pressure vapor state to a high-pressure vapor state or to a supercritical state before injecting it into the high-pressure loop 904. The location of this reinjection of excess flow is similarly different compared to that in FIG. 22A. Refrigerant in a vapor or supercritical state exiting the leak compressor 925 is injected downstream of the high-pressure outlet 922 of the pressure transducer 902 (which is at the same pressure as the leak compressor outlet pressure). As shown in FIG. 22B, a three-way valve 927 is located downstream of the evaporator 910 to allow splitting of the excess flow from the bulk flow in the low-pressure loop 906 before sending it through the leak compressor 925. Similarly, a three-way valve 929 is located downstream of the pressure transducer 902 to allow recombination of the high-pressure leak flow exiting the leak compressor 925 with the high-pressure bulk flow exiting the pressure transducer 902. This combined high-pressure flow then proceeds to the gas cooler / condenser 908, as previously described. The advantage of this configuration compared to that in Figure 22A is that it provides additional heat absorption capacity to the cycle due to the additional flow through the evaporator 910 (excess flow coming from the low pressure outlet 920). On the other hand, the energy consumption of this cycle is likely to be slightly higher than that of the system 900 shown in Figure 22A because the energy consumed by the leakage compressor 925 is slightly higher than the energy consumed by the multiphase leakage pump 913.This is because the refrigerant is compressed to a high pressure in a fully vapor state in the leak compressor 925, as opposed to being pumped in a partial or fully liquid state in the multiphase circulation pump 913.

[0087] The thermodynamic processes occurring within the refrigeration system 923 are described in further detail with reference to Figures 23 and 24. Figures 23 and 24 illustrate a temperature-entropy (TS) diagram 926 and a pressure-enthalpy (PH) diagram 928, respectively, to show the thermodynamic processes occurring in the four major components of the refrigeration system 900. Point 1 represents the leak compressor inlet 930 (see Figure 22B). Point 2 represents the leak compressor outlet 932 and the gas cooler inlet 934. Point 3 represents the gas cooler outlet 936 and the high-pressure inlet 914 of the rotary pressure exchanger 902. Point 4 represents the low-pressure outlet 920 of the rotary pressure exchanger 902 and the evaporator inlet 938. As illustrated in Figures 23 and 24, the leak compressor 925 raises the pressure and therefore the temperature of the refrigerant working fluid (e.g., carbon dioxide) above the environment, where it can reject heat to the hotter external environment. This occurs inside the gas cooler 908. In the gas cooler 908 of a supercritical carbon dioxide system, because the carbon dioxide is in a supercritical state, there is no phase boundary and the carbon dioxide is above the two-phase dome 940. Therefore, the temperature drops as the carbon dioxide rejects heat to the environment. As illustrated in Figures 23 and 24, the carbon dioxide in a supercritical state at the gas cooler outlet 936 enters the rotary pressure exchanger 902 at the high-pressure inlet port 914, undergoes isentropic or near-isentropic (approximately 85 percent isentropic efficiency) expansion, and exits the rotary pressure exchanger 902 at the low-pressure outlet port 920 as a two-phase gas-liquid carbon dioxide mixture. The two-phase carbon dioxide at point 4 then absorbs heat in the evaporator 910 (process 4 → 1, a constant enthalpy process). Overall, diagrams 926, 928 illustrate the cycle efficiency benefits resulting from increased cooling capacity and reduced compressor workload. Because the expansion inside the rotary pressure exchanger 902 occurs isentropically, it creates an enthalpy change that can be used to compress the fluid coming from the evaporator 910 to the total pressure in the system 900. This significantly reduces any work that would have been done by the bulk flow compressor, thus allowing its replacement by the leak compressor 925 (which consumes significantly less energy).

[0088] FIG. 25 is a schematic diagram of a refrigeration system 931 that uses low-DP recycle compressors instead of recycle pumps. The recycle compressors maintain fluid flow throughout the system 900, thereby minimizing pressure loss within the system 931. The difference between this system and the systems 900, 923 shown in FIGS. 22A and 22B is that the circulation of bulk flow within the low-pressure loop 906 and the high-pressure loop 904 is achieved using low-DP recycle compressors instead of low-DP multiphase recycle pumps. Similarly, the locations of these recycle compressors differ. For example, recycle compressor 941 (Compressor 1) in the low-pressure loop 906 is located downstream of evaporator 910, which circulates refrigerant in a vapor state. Similarly, recycle compressor 944 (Compressor 2) in the high-pressure loop 904 is located downstream of the high-pressure outlet 922 of pressure transducer 902, which circulates refrigerant in a supercritical or high-pressure vapor state. Compressor 3 is similar to the high-DP, low-flow leakage compressor 925 described in connection with FIG. 22B in that compressor 925 takes excess flow entering low-pressure loop 906 from pressure transducer 902 in a vapor state (e.g., leakage flow from pressure transducer 902) and compresses it back into high-pressure loop 904 as a high-pressure vapor or supercritical state. This excess flow is then combined with the high-pressure bulk flow from compressor 944 before proceeding to gas cooler / condenser 908. A low-DP recycle compressor 941 located along second fluid loop 906 (e.g., the low-pressure fluid loop) maintains fluid flow along loop 906 (e.g., between rotary pressure exchanger 902 and gas cooler 908). Additionally, a low-DP recycle compressor 944 located along first fluid loop 904 (e.g., the high-pressure fluid loop) maintains fluid flow along loop 904 (e.g., between evaporator 910 and rotary pressure exchanger 902). In some embodiments, refrigeration system 931 may include only compressors 925 and 941. In some embodiments, refrigeration system 900 may include only compressors 944 and 941. In some embodiments, compressors 941, 944 each have a significantly lower differential pressure across them than leak compressor 925, as noted in more detail below.

[0089] In some embodiments, a three-way valve is located at the junction between the flows exiting the compressors 925, 944 (e.g., near circle 2 in FIG. 25 ). This three-way valve is located in the high-pressure branch 904 between the high-pressure, high-flow, low-DP recycle compressor 944 and the gas cooler or condenser 908, where, during operation of the refrigeration system 931, a first flow from the high-DP, low-flow leakage compressor 925 is combined with the bulk flow exiting the high-pressure, high-flow, low-DP recycle compressor 944 before proceeding to the inlet 934 of the gas cooler or condenser 908. The high-pressure, high-flow, low-DP recycle compressor 944 is located between the high-pressure outlet 922 of the rotary pressure exchanger 902 and this three-way valve.

[0090] Similarly, in some embodiments, another three-way valve is located at the downstream junction (e.g., near the circled 1 in FIG. 25 ) of the evaporator 910 where it branches toward the compressors 925, 941. This three-way valve is located between the evaporator 910 and the rotary pressure exchanger 902 in the low-pressure branch 906, such that during operation of the refrigeration system 931, a portion of the flow exiting the evaporator 910 is diverted through the three-way valve to the inlet of the high-DP, low-flow leakage compressor 925, and the remaining portion of the flow passes to the low-pressure inlet 918 of the rotary pressure exchanger 902. A low-pressure, high-flow, low-DP recycle compressor is located between this three-way valve and the low-pressure inlet of the rotary pressure exchanger 902.

[0091] In a conventional refrigeration system (i.e., a supercritical carbon dioxide refrigeration system), the bulk flow compressor operates at a flow rate of approximately 30 gallons per minute and a differential pressure of approximately 10,342 kPa (1,500 psi). Assuming these operating conditions, the bulk flow compressor would require approximately 45,000 (i.e., 30 x 1,500 psi) units of force (i.e., work done or energy consumed). In the refrigeration system 900 described above, low DP recycle compressor 941 and low DP recycle compressor 944 (assuming each operates at a flow rate of approximately 30 gallons per minute and a differential pressure of approximately 10 psi) would each require approximately 300 (i.e., 30 x 10) units of force. Leak compressor 925 (assuming it operates at a flow rate of approximately 1.5 gallons and a differential pressure of approximately 1,500 psi) would require approximately 2,250 (i.e., 1.5 x 1,500) units of force. Compressors 925, 941, 944 in refrigeration system 931 would therefore require approximately 2,850 units of force. Compressors 925, 941, 944 would therefore reduce energy consumption by at least a factor of 10 (and even up to a factor of 15) compared to bulk flow compressor-based systems.

[0092] In some embodiments, the refrigeration system 931 (having a leakage compressor 925 and one or more of low DP circulation compressors 941, 944) may be used in the supermarket architecture described above in FIGS.

[0093] Figures 26 and 27 show two examples of supermarket system architectures 950, 952 using a rotary pressure exchanger-based supercritical carbon dioxide refrigeration system that also uses a conventional Joule-Thomson expansion valve 954. Generally, the architectures are similar to those in Figures 18 and 19, except for the use of the expansion valve 954. Additionally, while the architectures 950, 952 are discussed in connection with using a gas cooler as the heat exchanger 324 for use with a supercritical refrigerant (e.g., carbon dioxide), these architectures 950, 952 can also be used with a condenser as the heat exchanger 324 for use with a subcritical refrigerant (e.g., carbon dioxide). In the first architecture 950 (Figure 26), the two-phase low-pressure effluent stream (e.g., a carbon dioxide gas / liquid mixture at a first intermediate pressure, such as 370 psi) from the rotary pressure exchanger 304 (via low-pressure outlet 305) passes through a flash tank 306, which separates the gas and liquid phases (both exiting the flash tank at, e.g., 370 psi). The carbon dioxide liquid phase is transported to low temperature (e.g., approximately −20° C.) and medium temperature (e.g., approximately −4° C.) heat loads / evaporators 308, 310 (e.g., the freezer and refrigerator sections of a supermarket, respectively), where the carbon dioxide liquid phase picks up heat and becomes superheated. Because it is a pure liquid phase rather than a two-phase gas / liquid, it has greater heat absorption (i.e., cooling) capacity. The carbon dioxide liquid phase enters the medium temperature evaporator 310 at, for example, 370 psi, while the carbon liquid phase enters the low temperature evaporator 308 at, for example, 180 psi after flowing through a flow control valve 312. The flow control valve 312 can adjust (e.g., in response to a control signal from a controller) the flow rate of liquid carbon dioxide to the evaporator 308. The superheated carbon dioxide vapor from freezer section 308 (at a low pressure of 180 psi) then proceeds to cold compressor 316 (where it exits at, for example, 370 psi) before recombining with the superheated carbon dioxide vapor from refrigerator section 310 (at, for example, 370 psi) and with the separated superheated gas phase carbon dioxide that was separated from the gas / liquid mixture in flash tank 306, which is at the same pressure.A control valve 318 (e.g., a flash gas control valve) may adjust (e.g., in response to a control signal from a controller) the flow rate of superheated gaseous carbon dioxide flowing from flash tank 306. This recombined superheated gaseous carbon dioxide then enters rotary pressure exchanger 304 at low-pressure inlet port 320 and is compressed to a second intermediate pressure (e.g., 500 psi). The superheated gaseous carbon dioxide exits rotary pressure exchanger 304 (via high-pressure outlet 322) and proceeds to intermediate-temperature compressor 330, where the superheated gaseous carbon dioxide is compressed to the highest pressure in the system (e.g., 1,300 psi) depending on system requirements and converted to supercritical carbon dioxide. The supercritical carbon dioxide then proceeds at the highest pressure to heat exchanger 324 (e.g., a gas cooler), where it rejects heat to the environment and cools. In some embodiments, heat exchanger 324 is a gas condenser used with subcritical carbon dioxide. From the gas cooler 324, the supercritical carbon dioxide (e.g., at 1,300 psi) flows through a high-pressure Joule-Thomson valve 954 where the supercritical carbon dioxide is converted to a carbon dioxide gas / liquid mixture (e.g., at a second intermediate pressure, e.g., 500 psi). The carbon dioxide gas / liquid mixture flows into the high-pressure inlet 326 of the rotary pressure exchanger 304.

[0094] Architecture 952 in Figure 27 varies slightly from architecture 950 in Figure 26. Specifically, as depicted in Figure 27, the carbon dioxide gas / liquid mixture (at a second intermediate pressure, e.g., 500 psi) flows into flash tank 306 for separation into pure carbon dioxide gas or vapor and liquid. Carbon dioxide gas from flash tank 306 flows into high-pressure inlet 326 of rotary pressure exchanger 304, while carbon dioxide liquid from the flash tank flows at low pressure into low-temperature and intermediate-temperature evaporators 308, 310. The two-phase gas-liquid CO mixture exiting low-pressure inlet 305 of pressure transducer 304 exits at the same pressure as intermediate-temperature evaporator 310 and combines with the fluid stream exiting intermediate-temperature evaporator 310 and low-temperature compressor 316 before entering low-pressure inlet 320 of pressure transducer 304. Similarly, flow control valve 314 is located upstream of intermediate-temperature evaporator 310.

[0095] While the invention is susceptible to various modifications and variations, specific embodiments have been shown by way of example in the drawings and are described herein. It should be understood, however, that the invention is not intended to be limited to the particular forms disclosed. Rather, the invention is intended to cover all modifications, equivalents, and alternative forms falling within the spirit and scope of the invention as defined by the following appended claims. The following are some embodiments of the present invention. [Aspect 1] In a refrigeration system, A high pressure branch for circulating refrigerant at high pressure through it; a gas cooler or condenser disposed along the high-pressure branch, the high-pressure branch configured to reject heat from the refrigerant at high pressure to an ambient environment through the gas cooler or condenser, the refrigerant at high pressure being in a supercritical or subcritical state; a second low pressure branch for circulating said refrigerant at low pressure therethrough; an evaporator disposed along the low-pressure branch, the low-pressure branch configured to absorb heat from the ambient environment through the evaporator into the refrigerant at a low pressure, the refrigerant at a low pressure being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; a compressor or pump configured to increase the pressure of the refrigerant from a low pressure to a high pressure; a rotary pressure exchanger fluidly coupled to the low-pressure branch and the high-pressure branch, configured to receive the refrigerant at a high pressure from the high-pressure branch and to receive the refrigerant at a low pressure from the low-pressure branch and to exchange pressure between the refrigerant at a high pressure and the refrigerant at a low pressure, wherein a first exit stream from the rotary pressure exchanger comprises the refrigerant at a high pressure in a supercritical or subcritical state, and a second exit stream from the rotary pressure exchanger comprises the refrigerant at a low pressure in the liquid state or a two-phase mixture of the liquid and vapor; Refrigeration system including. [Aspect 2] 2. The refrigeration system of claim 1, wherein the refrigerant comprises carbon dioxide. [Aspect 3] 3. The refrigeration system of claim 1 or 2, wherein the rotary pressure exchanger is configured to enable the received refrigerant at a low pressure in the vapor state or the two-phase mixture of liquid and vapor to be compressed to a high-pressure refrigerant in a supercritical state or a subcritical state, and to enable the received refrigerant at a high pressure in the supercritical state or the subcritical state to be expanded to a low-pressure refrigerant in the two-phase mixture of liquid and vapor or the liquid state. [Aspect 4] 4. The refrigeration system of claim 3, wherein the evaporator is disposed downstream from the rotary pressure exchanger and configured to receive the refrigerant at a low pressure, the refrigerant being a two-phase mixture of liquid and vapor, and to convert the two-phase mixture of liquid and vapor into saturated vapor or superheated vapor. [Aspect 5] 5. A refrigeration system according to any one of the preceding aspects, comprising the compressor fluidly coupled to the low pressure and high pressure branches. [Aspect 6] A refrigeration system as described in aspect 5, wherein the evaporator is configured to provide a first portion of the low-pressure refrigerant in the vapor state to the rotary pressure exchanger and a second portion of the low-pressure refrigerant in the vapor state to the compressor, and the first and second portions of the low-pressure refrigerant in the vapor state include superheated vapor. [Aspect 7] 7. The refrigeration system of any one of claims 1 to 6, wherein the rotary pressure exchanger is configured to expand the high-pressure refrigerant in the supercritical state via isentropic or quasi-isentropic expansion to the low-pressure refrigerant that is a two-phase mixture of liquid and vapor. [Aspect 8] 8. The refrigeration system of any one of aspects 1 to 7, wherein the rotary pressure exchanger is used in place of a Joule-Thomson expansion valve to increase the cooling capacity of the refrigeration system and reduce the work requirements of the compressor. [Aspect 9] In a refrigeration system, A high pressure branch for circulating refrigerant at high pressure through it; a gas cooler or condenser disposed along the high-pressure branch, the high-pressure branch configured to reject heat from the refrigerant at high pressure to an ambient environment through the gas cooler or condenser, the refrigerant at high pressure being in a supercritical or subcritical state; A low-pressure branch for circulating refrigerant at low pressure through it; a first evaporator disposed along the low-pressure branch, the first evaporator configured to operate at a first temperature, the low-pressure branch configured to absorb heat from the ambient environment through an evaporator into the refrigerant at a low pressure, the refrigerant at a low pressure being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; a first intermediate pressure branch for circulating the refrigerant therethrough at a first intermediate pressure; a second evaporator disposed along the first intermediate pressure branch, the second evaporator configured to operate at a second temperature higher than the first temperature; a second intermediate pressure branch for circulating the refrigerant therethrough at a second intermediate pressure, wherein a first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the respective pressures of the refrigerant in the low pressure branch and the second intermediate pressure branch, the first intermediate pressure of the refrigerant in the first intermediate pressure branch is equal to a saturation pressure in the second evaporator, and the second intermediate pressure of the refrigerant in the second intermediate pressure branch is between the respective pressures of the refrigerant in the high pressure branch and the first intermediate pressure branch; a flash tank operating at the second intermediate pressure and configured to separate the two-phase liquid-vapor mixture of the refrigerant into a pure liquid and a pure vapor; a rotary pressure exchanger fluidly coupled to the second intermediate pressure branch and the high pressure branch, configured to receive the refrigerant at a high pressure from the high pressure branch and the refrigerant at a second intermediate pressure from the second intermediate pressure branch, the refrigerant being in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor, and to exchange pressure between the refrigerant at the high pressure and the refrigerant at the second intermediate pressure, wherein a first exit stream from the rotary pressure exchanger comprises the high pressure refrigerant in the supercritical state or the subcritical state, and a second exit stream from the rotary pressure exchanger comprises the refrigerant at the second intermediate pressure being in the liquid state or a two-phase mixture of the liquid and vapor; Refrigeration system including. [Aspect 10] 10. The refrigeration system of claim 9, further comprising: a first compressor positioned downstream of the flash tank and the first evaporator, the first compressor operating at a first temperature; and the first compressor configured to receive the refrigerant in the vapor state or the two-phase mixture of liquid and vapor from the first evaporator and to pressurize the refrigerant to the first intermediate pressure. [Aspect 11] 11. The refrigeration system of claim 10, further comprising: a second compressor positioned downstream of the first compressor and the second evaporator, the second compressor operating at a second temperature, the second compressor configured to receive refrigerant in the vapor state or the two-phase mixture of liquid and vapor from both the first compressor and the second evaporator, and to pressurize the refrigerant to the high pressure. [Aspect 12] 12. The refrigeration system of any one of aspects 9 to 11, comprising a first valve configured to adjust a flow rate of the separated liquid refrigerant from the flash tank to the first evaporator after the separated liquid refrigerant reaches the low pressure. [Aspect 13] 13. The refrigeration system of claim 12, further comprising: a second valve configured to adjust a flow rate of the separated liquid refrigerant from the flash tank to the second evaporator after the separated liquid refrigerant reaches the first intermediate pressure. [Aspect 14] 14. The refrigeration system of claim 13, comprising a third valve configured to regulate a flow of separated vapor refrigerant from the flash tank at the second intermediate pressure to an inlet of the rotary pressure exchanger. [Aspect 15] 15. The refrigeration system of any one of aspects 9 to 14, wherein the refrigerant comprises carbon dioxide. [Aspect 16] In a refrigeration system, A high pressure branch for circulating refrigerant at high pressure through it; a gas cooler or condenser disposed along the high-pressure branch, the high-pressure branch configured to reject heat from the refrigerant at high pressure to an ambient environment through the gas cooler or condenser, the refrigerant at high pressure being in a supercritical or subcritical state; A low-pressure branch for circulating refrigerant at low pressure through it; a first evaporator disposed along the low-pressure branch, the first evaporator configured to operate at a first temperature, the low-pressure branch configured to absorb heat from the ambient environment through an evaporator into the refrigerant at a low pressure, the refrigerant at a low pressure being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; an intermediate pressure branch for circulating said refrigerant at an intermediate pressure therethrough; a second evaporator disposed along the intermediate-pressure branch, the second evaporator configured to operate at a second temperature higher than the first temperature, wherein an intermediate pressure of the refrigerant in the intermediate-pressure branch is between the respective pressures of the refrigerant in the high-pressure branch and the low-pressure branch, and the intermediate pressure of the refrigerant in the intermediate-pressure branch is equal to a saturation pressure in the second evaporator; a flash tank operating at the intermediate pressure and configured to separate the two-phase liquid-vapor mixture of the refrigerant into a pure liquid and a pure vapor; a rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high pressure branch, configured to receive the refrigerant at a high pressure from the high pressure branch and to receive the refrigerant at an intermediate pressure from the intermediate pressure branch, the refrigerant being in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor, and to exchange pressure between the refrigerant at the high pressure and the refrigerant at the intermediate pressure, wherein a first exit stream from the rotary pressure exchanger comprises the refrigerant at a high pressure in a supercritical state or a subcritical state, and a second exit stream from the rotary pressure exchanger comprises the refrigerant at an intermediate pressure in the liquid state or a two-phase mixture of the liquid and vapor; Refrigeration system including. [Aspect 17] 17. The refrigeration system of claim 16, comprising a low differential pressure compressor configured to receive the refrigerant in the supercritical or subcritical state exiting the rotary pressure exchanger and to compress the refrigerant to the high pressure. [Aspect 18] 18. The refrigeration system of claim 17, further comprising: a compressor positioned downstream of the flash tank and the first evaporator, the first compressor operating at the first temperature, the compressor configured to receive the refrigerant in the vapor state or the two-phase mixture of liquid and vapor from the first evaporator, and to pressurize the refrigerant to the intermediate pressure for the rotary pressure exchanger. [Aspect 19] 20. The refrigeration system of claim 18, comprising a valve configured to regulate a flow of separated vapor refrigerant from the flash tank at the intermediate pressure to an inlet of the rotary pressure exchanger. [Aspect 20] 20. The refrigeration system of any one of aspects 16 to 19, wherein the refrigerant comprises carbon dioxide.

Claims

1. In a refrigeration system, a high pressure branch for circulating refrigerant at high pressure therethrough; a gas cooler or condenser disposed along the high-pressure branch, the high-pressure branch configured to release first heat from the refrigerant at high pressure through the gas cooler or condenser to a first ambient environment, the refrigerant at high pressure being in a supercritical or subcritical state; a low pressure branch for circulating refrigerant at low pressure therethrough; a first evaporator disposed along the low-pressure branch, the first evaporator configured to operate at a first temperature, the low-pressure branch configured to absorb second heat from a second ambient environment through the first evaporator into the refrigerant at a low pressure, the refrigerant at a low pressure being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; a first intermediate pressure branch for circulating the refrigerant therethrough at a first intermediate pressure; a second evaporator disposed along the first intermediate pressure branch, the second evaporator configured to operate at a second temperature higher than the first temperature; a second intermediate pressure branch for circulating the refrigerant therethrough at a second intermediate pressure, wherein a first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the respective pressures of the refrigerant in the low pressure branch and the refrigerant in the second intermediate pressure branch, the first intermediate pressure of the refrigerant in the first intermediate pressure branch is equal to a saturation pressure in the second evaporator, and the second intermediate pressure of the refrigerant in the second intermediate pressure branch is between the respective pressures of the refrigerant in the high pressure branch and the refrigerant in the first intermediate pressure branch; a flash tank operating at the second intermediate pressure and configured to separate the two-phase liquid-vapor mixture of the refrigerant into a substantially pure liquid and a pure vapor; a rotary pressure exchanger fluidly coupled to the second intermediate pressure branch and the high pressure branch, the rotary pressure exchanger comprising: receiving the refrigerant at high pressure from the high-pressure branch; receiving the refrigerant at the second intermediate pressure from the second intermediate pressure branch, the refrigerant being in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor; and a rotary pressure exchanger configured to exchange pressure between the high pressure refrigerant and the second intermediate pressure refrigerant, wherein a first exit stream from the rotary pressure exchanger comprises the high pressure refrigerant in the supercritical state or the subcritical state, and a second exit stream from the rotary pressure exchanger comprises the refrigerant at the second intermediate pressure in the liquid state or as a two-phase mixture of the liquid and vapor; Refrigeration system including.

2. 2. The refrigeration system of claim 1, further comprising a first compressor positioned downstream of the flash tank and the first evaporator, the first compressor configured to operate at a first temperature, the first compressor configured to receive the refrigerant in the vapor state or the two-phase mixture of liquid and vapor from the first evaporator, and to pressurize the refrigerant to the first intermediate pressure.

3. 3. The refrigeration system of claim 2, further comprising a second compressor positioned downstream of the first compressor and the second evaporator, the second compressor operating at a second temperature, the second compressor configured to receive refrigerant in the vapor state or the two-phase mixture of liquid and vapor from both the first compressor and the second evaporator, and to pressurize the refrigerant to the high pressure.

4. 2. The refrigeration system of claim 1, further comprising a first valve configured to adjust a first flow rate of the separated liquid refrigerant from the flash tank so that the separated liquid refrigerant flows to the first evaporator after reaching the low pressure.

5. 5. The refrigeration system of claim 4, further comprising a second valve configured to regulate a second flow rate of the separated liquid refrigerant from the flash tank to the second evaporator after the separated liquid refrigerant reaches the first intermediate pressure.

6. 6. The refrigeration system of claim 5, further comprising a third valve configured to regulate a third flow rate of separated vapor refrigerant from the flash tank at the second intermediate pressure to an inlet of the rotary pressure exchanger.

7. The refrigeration system of claim 1 , wherein the refrigerant comprises carbon dioxide.

8. In a refrigeration system, a high pressure branch for circulating refrigerant at high pressure therethrough; a gas cooler or condenser disposed along the high-pressure branch, the high-pressure branch configured to release first heat from the refrigerant at high pressure through the gas cooler or condenser to a first ambient environment, the refrigerant at high pressure being in a supercritical or subcritical state; a low pressure branch for circulating refrigerant at low pressure therethrough; a first evaporator disposed along the low-pressure branch, the first evaporator configured to operate at a first temperature, the low-pressure branch configured to absorb second heat from a second ambient environment through the first evaporator into the refrigerant at a low pressure, the refrigerant at a low pressure being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; an intermediate pressure branch for circulating said refrigerant at an intermediate pressure therethrough; a second evaporator disposed along the intermediate-pressure branch, the second evaporator configured to operate at a second temperature higher than the first temperature, wherein an intermediate pressure of the refrigerant in the intermediate-pressure branch is between respective pressures of the refrigerant in the high-pressure branch and the low-pressure branch, the intermediate pressure of the refrigerant in the intermediate-pressure branch being substantially equal to a saturation pressure in the second evaporator; a flash tank operating at the intermediate pressure and configured to separate the two-phase liquid-vapor mixture of the refrigerant into a substantially pure liquid and a pure vapor; a rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high pressure branch, the rotary pressure exchanger comprising: receiving the refrigerant at high pressure from the high-pressure branch; receiving the refrigerant at the intermediate pressure from the intermediate pressure branch, in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor; a rotary pressure exchanger configured to exchange pressure between the refrigerant at a high pressure and the refrigerant at an intermediate pressure, wherein a first exit stream from the rotary pressure exchanger contains the refrigerant at a high pressure in a supercritical or subcritical state, and a second exit stream from the rotary pressure exchanger contains the refrigerant at an intermediate pressure in the liquid state or as a two-phase mixture of the liquid and vapor; Refrigeration system including.

9. 9. The refrigeration system of claim 8, further comprising a low differential pressure compressor configured to receive the refrigerant in the supercritical or subcritical state exiting the rotary pressure exchanger and compress the refrigerant to the high pressure.

10. 18. The refrigeration system of claim 17, further comprising a second compressor positioned downstream of the flash tank and the first evaporator, the second compressor configured to operate at the first temperature, the second compressor configured to receive the refrigerant in the vapor state or the two-phase mixture of liquid and vapor from the first evaporator, and to pressurize the refrigerant to the intermediate pressure for the rotary pressure exchanger.

11. The refrigeration system of claim 10 , further comprising a valve configured to regulate the flow of separated vapor refrigerant from the flash tank at the intermediate pressure to an inlet of the rotary pressure exchanger.

12. The refrigeration system of claim 8 , wherein the refrigerant comprises carbon dioxide.

13. circulating refrigerant at high pressure to the high pressure branch; releasing a first heat from the refrigerant at high pressure to a first ambient environment via a gas cooler or condenser disposed along the high pressure branch, wherein the refrigerant at high pressure is in a supercritical or subcritical state; circulating refrigerant at low pressure to the low pressure branch; absorbing second heat into the refrigerant from a second ambient environment via a first evaporator disposed along the low-pressure branch, the first evaporator configured to operate at a first temperature, and the refrigerant at low pressure is in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; circulating the refrigerant at an intermediate pressure to the intermediate pressure branch; absorbing third heat into the refrigerant from a third ambient environment via a second evaporator disposed along the intermediate pressure branch, the second evaporator configured to operate at a second temperature greater than the first heat, an intermediate pressure of the refrigerant in the intermediate pressure branch being between respective pressures of the refrigerant in the high pressure branch and the refrigerant in the low pressure branch, the intermediate pressure of the refrigerant in the intermediate pressure branch being substantially equal to a saturation pressure in the second evaporator; separating the refrigerant operating at the intermediate pressure and being a two-phase mixture of liquid and vapor via a flash tank into a substantially pure liquid and a pure vapor; receiving the refrigerant at high pressure from the high pressure branch via a rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high pressure branch; receiving the refrigerant at the intermediate pressure through the rotary pressure exchanger in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor; exchanging pressure between the refrigerant at a high pressure and the refrigerant at an intermediate pressure through the rotary pressure exchanger, wherein a first exit stream from the rotary pressure exchanger contains the refrigerant at a high pressure in a supercritical or subcritical state, and a second exit stream from the rotary pressure exchanger contains the refrigerant at an intermediate pressure in the liquid state or a two-phase mixture of the liquid and vapor; A method comprising:

14. receiving the refrigerant in the supercritical or subcritical state exiting the rotary pressure exchanger via a low differential pressure compressor; The method of claim 13 further comprising compressing the refrigerant to the high pressure through the low differential pressure compressor.

15. receiving the refrigerant in the vapor state or the two-phase mixture of liquid and vapor from the first evaporator through a second compressor positioned downstream of the flash tank and the first evaporator, the second compressor configured to operate at the first temperature; 15. The method of claim 14, further comprising compressing the refrigerant to the intermediate pressure for the rotary pressure exchanger via the second compressor.

16. 16. The method of claim 15, further comprising regulating the flow of separated vapor refrigerant from the flash tank at the intermediate pressure to an inlet of the rotary pressure exchanger via a valve.

17. The method of claim 13 , wherein the refrigerant comprises carbon dioxide.

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