Refrigeration system with high-speed rotating 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 the technical field that may be relevant to the various aspects of the invention described and / or claimed below. This discussion is believed to be useful in providing the reader with background information to facilitate a better understanding of the various aspects of the invention. Accordingly, these statements should be read from that perspective and it should be understood that they are not an admission of prior art.

Background Art

[0002] With the enforcement of 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 Europe and much of North America, they face drawbacks in hot climates because their coefficient of performance (measure of efficiency) decreases as the ambient temperature of the surrounding environment increases, resulting in higher electricity costs per unit of cooling performed. This is because supercritical carbon dioxide systems require much higher pressures to operate (approximately 10,342 kPa (1500 psi) or more) compared to HFC / CFC-based systems (approximately 1,379–2,068.4 kPa (200–300 psi)). Very high differential pressure compressors are used to raise the refrigerant above critical pressure. A larger pressure ratio across the compressor consumes more electrical energy. This problem is more pronounced in hotter climates because the refrigerant temperature at the inlet of the cooling system needs to be raised to a sufficiently high temperature to allow heat to be released into the hotter ambient environment. This is done by further increasing the pressure ratio across the compressor, thus requiring more electricity from the compressor, and subsequently increasing the electricity cost 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, increase their availability, and simultaneously help mitigate global warming. [Overview of the Initiative]

[0003] Several embodiments consistent with the subject matter and scope of the disclosure are summarized below. These embodiments are not intended to limit the scope of the disclosure, but rather to provide a brief summary of some of the disclosed embodiments. In fact, the disclosure may encompass a variety of forms that may be similar to or different from the embodiments described below.

[0004] In one embodiment, a refrigeration system is provided. The refrigeration system includes a high-pressure branch for circulating a refrigerant at high pressure through it. The refrigeration system also includes a gas cooler or condenser located along the high-pressure branch, where the high-pressure branch is configured to release heat from the refrigerant to the ambient environment at high pressure through this gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state. The refrigeration system further includes a low-pressure branch for circulating a refrigerant at low pressure through it. The refrigeration system further also includes an evaporator located along the low-pressure branch, where the low-pressure branch is configured to absorb heat from the ambient environment into the low-pressure refrigerant through this evaporator, and the low-pressure refrigerant is in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. The refrigeration system further also includes a compressor or pump configured to increase the pressure of the refrigerant from low pressure to high pressure. The refrigeration system further includes a rotary pressure exchanger fluid-coupled to a low-pressure branch and a high-pressure branch, the rotary pressure exchanger being configured to receive high-pressure refrigerant from the high-pressure branch and low-pressure refrigerant from the low-pressure branch, and to exchange pressure between the high-pressure and low-pressure refrigerants, wherein the first outflow from the rotary pressure exchanger contains high-pressure refrigerant in a supercritical or subcritical state, and the second outflow from the rotary pressure exchanger contains low-pressure refrigerant in a liquid state or as 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 high pressure through a medium. The refrigeration system includes a gas cooler or condenser located along the high-pressure branch, the high-pressure branch being configured to release heat from the refrigerant to the ambient environment at high pressure through the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state. The refrigeration system similarly includes a low-pressure branch for circulating a refrigerant at low pressure through a medium. The refrigeration system further also includes a first evaporator located along the low-pressure branch, the first evaporator being configured to operate at a first temperature, and the low-pressure branch being configured to absorb heat from the ambient environment into the low-pressure refrigerant through the evaporator, and the low-pressure refrigerant is 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 through a medium at a first intermediate pressure. The refrigeration system further includes a second evaporator located along a first intermediate pressure branch, where this 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 through it at a second intermediate pressure, where the first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the respective pressures of the low-pressure branch and the second intermediate pressure branch, where the first intermediate pressure of the refrigerant in the first intermediate pressure branch is equal to the 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 high-pressure branch and the first intermediate pressure branch.The refrigeration system further includes a flash tank configured to operate at a second intermediate pressure and separate a two-phase mixture of liquid and vapor refrigerant into pure liquid and pure vapor; and a rotary pressure exchanger fluid-coupled to a second intermediate pressure branch and a high-pressure branch, the rotary pressure exchanger configured to receive high-pressure refrigerant from the high-pressure branch and a second intermediate-pressure refrigerant from the second intermediate-pressure branch in vapor state, liquid state, or two-phase mixture of liquid and vapor, and to exchange pressure between the high-pressure refrigerant and the second intermediate-pressure refrigerant, wherein a first outflow from the rotary pressure exchanger contains high-pressure refrigerant in a supercritical or subcritical state, and a second outflow from the rotary pressure exchanger contains a second intermediate-pressure refrigerant in liquid state or 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 high pressure through a medium. The refrigeration system also includes a gas cooler or condenser located along the high-pressure branch, where the high-pressure branch is configured to release heat from the refrigerant to the ambient environment at high pressure through the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state. The refrigeration system further includes a second low-pressure branch for circulating a refrigerant at low pressure through a medium. The refrigeration system further also includes a first evaporator located along the low-pressure branch, where this first evaporator is configured to operate at a first temperature, where the low-pressure branch is configured to absorb heat from the ambient environment into the low-pressure refrigerant through this evaporator, and the low-pressure refrigerant is in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor. The refrigeration system further also includes an intermediate-pressure branch for circulating a refrigerant at an intermediate pressure through a medium. The refrigeration system further includes a second evaporator located along an intermediate pressure branch, which is configured to operate at a second temperature higher than the first temperature. The 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 the saturation pressure in the second evaporator. The refrigeration system further includes a flash tank which operates at the intermediate pressure and is configured to separate the two-phase mixture of liquid and vapor refrigerant into pure liquid and pure vapor. The refrigeration system further includes a rotary pressure exchanger fluid-coupled to an intermediate pressure branch and a high pressure branch, the rotary pressure exchanger being configured to receive high-pressure refrigerant from the high-pressure branch and intermediate-pressure refrigerant from the intermediate-pressure branch in vapor state, liquid state, or a two-phase mixture of liquid and vapor, and to exchange pressure between the high-pressure refrigerant and the intermediate-pressure refrigerant, wherein the first outflow from the rotary pressure exchanger contains high-pressure refrigerant in a supercritical or subcritical state, and the second outflow from the rotary pressure exchanger contains 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 be better understood by reading the following detailed description while referring to the accompanying drawings, in which similar letters throughout the figures represent similar parts. [Brief explanation of the drawing]

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

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

[0010] [Figure 3] Figure 3 is a temperature-entropy diagram showing the thermodynamic process in the refrigeration system using a Joule-Thomson expansion valve versus the refrigeration system in Figure 2.

[0011] [Figure 4] Figure 4 shows the pressure-enthalpy diagram of the thermodynamic process in the refrigeration system using a Joule-Thomson expansion valve versus the refrigeration system in Figure 2.

[0012] [Figure 5] Figure 5 is a disassembled and assembled perspective view of one embodiment of a rotary pressure exchanger or rotary LPC.

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

[0014] [Figure 7] Figure 7 is an exploded and assembled 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 the third operating position.

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

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

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

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

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

[0021] [Figure 14] FIG. 14 is a cross-sectional view of one embodiment of a barrier along line 14-14 of FIG. 11.

[0022] [Figure 15] FIG. 15 is a cross-sectional view of one embodiment of a barrier along line 14-14 of FIG. 11.

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

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

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

[0027] [Figure 20] Figure 20 is a schematic diagram of one embodiment of a control system that controls the movement of the power fluid and working fluid within the RLPC.

[0028] [Figure 21] Figure 21 is a schematic diagram of one embodiment of a control system for controlling the movement of the power fluid and working fluid within the RLPC.

[0029] [Figure 22A] Figure 22A is a schematic diagram of one embodiment of a refrigeration system having a rotary pressure exchanger or rotary liquid piston compressor (LPC) (for example, one having a low-flow, high-differential-pressure (DP) leak pump and a low-DP, high-flow circulation pump instead of a bulk-flow compressor).

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

[0031] [Figure 23] Figure 23 is the temperature-entropy diagram of the thermodynamic process in the refrigeration system shown in Figure 22.

[0032] [Figure 24]Figure 24 is a pressure-enthalpy diagram of the thermodynamic processes within the refrigeration system shown in Figure 22.

[0033] [Figure 25] Figure 25 is a schematic diagram of one embodiment of a refrigeration system having a rotary pressure exchanger or rotary liquid piston compressor (LPC) (for example, having a leak compressor instead of a bulk flow compressor and an additional low-DP circulating compressor (e.g., a blower)).

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

[0035] [Figure 27] Figure 27 is a schematic diagram of one embodiment of a refrigeration system in an alternative supermarket refrigeration system architecture (for example, having an expansion valve). [Modes for carrying out 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. Furthermore, in order to provide a concise description of these exemplary embodiments, not all features of actual implementations may be described herein. As with any engineering or design project, it should be recognized that in developing any such actual implementation, many implementation-specific decisions must be made to achieve the developer's specific end objectives, such as compatibility with system-related and business-related constraints, which may vary from implementation to implementation. Moreover, it should be recognized that although such development efforts may be complex and time-consuming, they will appear to be routine design, fabrication, and manufacturing work for those skilled in the art who benefit from this disclosure. Furthermore, it should be recognized that the features of the different embodiments disclosed herein are combined with each other unless otherwise noted.

[0037] The following discussion describes refrigeration systems (e.g., supercritical carbon dioxide refrigeration systems) that utilize a rotary pressure exchanger or a 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 its cooling capacity while recovering a large portion of the pressure energy that would have been lost if a Joule-Thomson expansion valve were used. Replacing the Joule-Thomson expansion valve with a rotary pressure exchanger increases efficiency due to the elimination of both energy destruction and entropy generation that occur within the expansion valve, resulting in total losses of up to 40 percent in a typical refrigeration system. Furthermore, replacing the Joule-Thomson expansion valve with a rotary pressure exchanger increases efficiency by changing the expansion process from an isenthalpy (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, the rotary pressure exchanger can also similarly replace the function of a bulk flow compressor. Thus, it becomes possible to use one or more low differential pressure (DP) circulating compressors (blowers) or circulating pumps in place of bulk flow high differential pressure compressors, and to maintain flow rates within the refrigeration system (for example, to overcome small pressure losses). These low DP circulating compressors may consume significantly less energy than bulk flow compressors (e.g., less than one-tenth). Replacing both Joule-Thomson expansion valves and bulk flow compressors with rotary pressure exchangers eliminates two of the largest sources of inefficiency in a refrigeration system while reducing power consumption and electricity costs. Furthermore, using rotary pressure exchangers in place of expansion valves and / or bulk flow compressors may increase the usability of the refrigeration system in other environments (e.g., warmer environments).At warmer ambient temperatures (e.g., 50 degrees Celsius), the compressor pressure ratio is altered (due to a significant increase in the pressure required at the compressor outlet), causing a dramatic 60 percent decrease in cycle efficiency (i.e., coefficient of performance) compared to the optimal temperature (e.g., 35 degrees Celsius). A rotating pressure exchanger mitigates the adverse effects of warmer ambient temperatures on the required compressor work, the cooling capacity of the refrigeration system, and the coefficient of performance of the refrigeration system.

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

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

[0040] Figure 2 is a schematic diagram of one embodiment of a refrigeration system 800 (e.g., a supercritical carbon dioxide refrigeration system) that uses a fluid in a supercritical state. Although the refrigeration system 800 is described as using carbon dioxide, other refrigerants are also available. By using a rotary pressure exchanger or rotary liquid compressor 802 (indicated as PX in the figure) as described below in place of an expansion valve (e.g., a Joule-Thomson expansion valve) within the refrigeration system 800, the refrigeration system 800 can operate more efficiently by increasing its cooling capacity while recovering a large portion of the pressure energy that would have been lost if a Joule-Thomson expansion valve had been used. In some embodiments, the rotary pressure exchanger can replace the function of a bulk flow compressor, thus allowing the use of one or more (significantly more energy-efficient) low-DP circulating compressors or pumps in place of the bulk flow compressor. For example, supercritical carbon dioxide refrigeration systems need to operate at much higher pressures (approximately 10,342 kPa (1500 psi) or more), which creates a large pressure ratio across the compressor (a compressor with a very high differential pressure), resulting in the consumption of more electrical energy. By replacing the expansion valve with a rotary pressure exchanger, it becomes possible to recover almost all of the pressure drop within the rotary pressure exchanger and use it to pressurize the flow coming from the evaporator rather than sending the flow to the main compressor. Thus, the electricity demand for the compressor can be significantly reduced or eliminated. Refrigeration systems 800, which use a rotary pressure exchanger instead of a Joule-Thomson expansion valve and / or bulk flow compressor, can be used in a wide range 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 combination of refrigeration and power generation cycles), aquariums, polar habitat research systems, and any other systems where refrigeration is used.

[0041] As described, the 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 rotating pressure exchanger 802. The heat exchanger 808 releases heat from the high-pressure refrigerant to the surrounding environment. Although the gas cooler is described below for use with a supercritical high-pressure refrigerant (e.g., carbon dioxide), in some embodiments the condenser can be used with a subcritical high-pressure refrigerant (e.g., carbon dioxide). The subcritical state for a refrigerant is below the critical point (more 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 a rotating pressure exchanger 802. The heat exchanger 810 absorbs heat from the surrounding environment into the low-pressure refrigerant. The low-pressure refrigerant in the low-pressure branch 806 may be in liquid state, vapor state, or a two-phase mixture of liquid and vapor. Both fluid loops 804 and 806 are fluid-coupled to a compressor 812 (e.g., a bulk-flow compressor). The compressor 812 converts the superheated gaseous carbon dioxide received from the evaporator 810 into supercritical carbon dioxide supplied to the gas cooler 808 (by increasing temperature and pressure). In some embodiments, as will be further detailed below, the compressor 812 may be replaced by one or more low-DP circulating compressors or pumps to overcome small pressure losses within the system 800 and maintain fluid flow. Generally, along the first fluid loop 804, the gas cooler 808 receives supercritical carbon dioxide, which is then cooled somewhat (e.g., at the high-pressure inlet 822) before being supplied to the rotating pressure exchanger 802. Along the second fluid loop 804, the evaporator 810 supplies a first portion of superheated gaseous carbon dioxide to the low-pressure inlet 813 of the rotary pressure exchanger 802 and a second portion of superheated gaseous carbon dioxide to the compressor 812. The rotary pressure exchanger 802 exchanges pressure between supercritical carbon dioxide and superheated gaseous carbon dioxide.Supercritical carbon dioxide is converted into a two-phase liquid / vapor mixture inside the rotating pressure exchanger 802 and exits through the low-pressure outlet 824, supplying it to the evaporator 810. The rotating pressure exchanger 802 similarly increases the pressure and temperature of superheated gaseous carbon dioxide, converting it into supercritical carbon dioxide, which exits the rotating pressure exchanger 802 via the high-pressure outlet 815, which supplies it to the gas cooler 808. As illustrated in Figure 2, the supercritical carbon dioxide exiting the rotating pressure exchanger 802 may be combined with carbon dioxide supplied from the compressor 812 to the gas cooler 808.

[0042] The thermodynamic processes occurring within the refrigeration system 800 (for example, in relation to a refrigeration system using a Joule-Thomson expansion valve) are described in more detail with reference to Figures 3 and 4. Figures 3 and 4 illustrate the temperature-entropy (TS) diagram 814 and the pressure-entropy (PH) diagram 816, respectively, to show the thermodynamic processes occurring in four main components of the refrigeration system 800 compared to a refrigeration system including a Joule-Thomson expansion valve. Point 1 represents the compressor inlet 818 (see Figure 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 (indicated as PX in Figures 3 and 4) and the evaporator inlet 826. As illustrated in Figures 3 and 4, the compressor 812 increases the pressure, thus raising the temperature of the refrigerant working fluid (e.g., carbon dioxide) to a temperature higher than the environment, where it can release heat into the warmer external environment. This occurs inside the gas cooler 808. Because the carbon dioxide is in a supercritical state, there is no phase boundary, unlike conventional condensers where the temperature remains constant for 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, and the carbon dioxide is above the two-phase dome 828. Therefore, the temperature decreases as the carbon dioxide releases heat into the environment. The higher the ambient temperature, the greater the pressure ratio across the compressor 812 and the greater the system pressure. At point 3, the carbon dioxide leaving the gas cooler outlet 830 then travels through the expansion valve (in a refrigeration system with a Joule-Thomson expansion valve) and follows a constant enthalpy process within the valve (3→4h), as shown by curve 832. On the PH diagram 816, curve 832 is a straight vertical line (because it is an isenthalpy process). As a result, 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 then continues through the evaporator 810, where the liquid carbon dioxide absorbs even more heat, becoming saturated vapor at the outlet 836 of the evaporator 810. Thus, the fluid entering the compressor 812 is in the pure vapor phase (gas phase).

[0043] Here, we consider a system having a rotary pressure exchanger 802 that replaces the Joule-Thomson expansion valve as shown in Figure 2. As illustrated in Figures 3 and 4, supercritical carbon dioxide at the gas cooler outlet 830 enters the rotary pressure exchanger 802 at the high-pressure inlet port 822, undergoes isentropic or near-isentropic expansion (e.g., 85 percent isentropic efficiency), and exits the rotary pressure exchanger 802 as two-phase gas-liquid carbon dioxide at the low-pressure outlet port 824. This process is represented by curve 835 on TS and PH diagrams 814 and 816. As illustrated, curve 835 (obtained using the rotary pressure exchanger 802) is located to the left of curve 832 (obtained using the expansion valve), which means that the amount or percentage of liquid content in the two-phase liquid is different in the case of the expansion valve (point 4 on PH diagram 816). h This means that the expansion through the rotating pressure exchanger 802 (position 4 on PH diagram 816) is greater than the expansion through the Joule-Thomson expansion valve (position 4). Due to the larger 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 rotating pressure exchanger 802 is used instead of the Joule-Thomson expansion valve. The position of point 4s on PH diagram 816 represents a complete isentropic expansion process (e.g., 100 percent isentropic expansion coefficient). The two-phase carbon dioxide at point 4 will then absorb heat in the evaporator 810 (process 4→1). Segment 840 length 838(4 hThe additional cooling capacity provided by the system 800 using the rotary pressure exchanger 802 is compared to a typical system 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 main advantages provided by incorporating the rotary pressure exchanger 802 into the refrigeration cycle.

[0044] Another advantage provided by using the rotating pressure exchanger 802 in the refrigeration cycle is that superheated gaseous carbon dioxide enters the rotating pressure exchanger 802 from the evaporator 80 (at the low-pressure inlet 813) and undergoes isentropic or near-isentropic compression (e.g., 85 percent isentropic efficiency) as shown by the dashed line 842 (i.e., process 1 → 2 s This becomes apparent when we observe the second fluid stream. This process is analogous to the isentropic process 1→2 that occurs inside the compressor 812. Since almost all of the compression occurs inside the rotating pressure exchanger 802, in some embodiments it is possible to completely or partially remove the main compressor 812. For example, in this case, the compressor 812 can be replaced by a gas blower or circulation pump with a very small differential pressure that consumes very little work (due to the very small enthalpy change across it). This results in a tremendous advantage to the efficiency of the refrigeration cycle, as can be seen from the following equation for performance efficiency (COP) (i.e., the standard measure of the efficiency of the refrigeration cycle):

number

[0045] Figure 5 is an exploded and assembled perspective view of one embodiment of a rotary pressure exchanger or rotary liquid piston compressor 40 (rotary LPC) (e.g., rotary pressure exchanger 802 in Figure 2) having the ability to transfer 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 fluid mixing. The rotary LPC 40 may include a substantially cylindrical body portion 42, which includes a sleeve 44 (e.g., rotor sleeve) and a rotor 46. The rotary LPC 40 may also include two end caps 48 and 50, which include manifolds 52 and 54, respectively. Manifold 52 includes inlet and outlet ports 56 and 58, respectively, while manifold 54 includes inlet and outlet ports 60 and 62, respectively. During operation, these inlet ports 56, 60 allow first and second fluids to enter the rotating LPC 40 and exchange pressure, while outlet ports 58, 62 allow the first and second fluids to exit the rotating LPC 40. During operation, inlet port 56 can receive the high-pressure first fluid, and after the pressure exchange, outlet port 58 can be used to guide the low-pressure first fluid out of the rotating LPC 40. Similarly, inlet port 60 can receive the low-pressure second fluid, and outlet port 62 can be used to guide the high-pressure second fluid out of the rotating LPC 40. The end caps 48 and 50 include respective end covers 64 and 66 located inside the respective manifolds 52 and 54, which enable fluid sealing contact with the rotor 46. The rotor 46 may be cylindrical and located within the sleeve 44, which allows the rotor 46 to rotate about an axis 68. The rotor 46 may have a plurality of channels 70 that extend substantially longitudinally through the rotor 46, with openings 72 and 74 at each end arranged symmetrically around the longitudinal axis 68. The openings 72 and 74 of the rotor 46 are arranged to fluidly communicate 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-pressure and low-pressure fluids during rotation.As illustrated, the inlet and outlet apertures 76 and 78; 80 and 82 may be designed in the form of an arc or a segment of a circle (e.g., C-shape).

[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 in the rotating LPC 40, thereby improving the operability of the fluid operating system. For example, by varying the volumetric flow rates of the first and second fluids entering the rotating LPC 40, a plant operator (e.g., a system operator) can control the amount of fluid mixing inside the rotating liquid piston compressor 10. Furthermore, by varying the rotational speed of the rotor 46, the operator can similarly control the mixing. Three characteristics of the rotating LPC 40 that affect the 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 fluid barrier (e.g., boundary portion) between the first and second fluids inside the rotor channel 70. Firstly, the rotor channels 70 are generally long and narrow, which stabilizes the flow rate inside the rotating LPC 40. Furthermore, the first and second fluids may move through the channels 70 in a plug flow regime with minimal axial mixing. Secondly, 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. Thirdly, only a small portion of the rotor channels 70 is used for pressure exchange between the first and second fluids. Thus, a certain volume of fluid remains in the channels 70 as a barrier between the first and second fluids. All of these mechanisms may limit mixing within the rotating LPC 40. Furthermore, in some embodiments, the rotating LPC40 may be designed to operate using a full or partial internal piston or other barrier that isolates the first and second fluids while allowing pressure transfer.

[0047] Figures 6-9 are exploded views of one embodiment of the rotating LPC 40 illustrating the sequence of positions of a single rotor channel 70 within the rotor 46 as the channel 70 completes one full cycle of rotation. Figures 6-9 are simplified representations of the rotating LPC 40 showing one rotor channel 70, the channel 70 being shown as having a circular cross-sectional shape. In other embodiments, the rotating LPC 40 may include multiple channels 70 having the same or different cross-sectional shapes (e.g., circular, signet, square, rectangular, polygonal, etc.). Thus, Figures 6-9 are simplified for illustrative purposes, and other embodiments of the rotating LPC 40 may have configurations different from those shown in Figures 6-9. As will be detailed below, the rotating LPC 40 facilitates pressure exchange between the first and second fluids by allowing the first and second fluids to come into brief contact with each other inside 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 velocity of the pressure wave traveling through the rotor channel 70 (immediately upon exposure of the channel to the aperture 76), the fluid diffusion rate, and the rotational speed of the rotor 46 determine whether and to what extent any mixing is occurring.

[0048] In Figure 6, the channel opening 72 is in a first position. In this first position, the channel opening 72 is in fluid communication with the aperture 78 in the end cover 64 and subsequently with the manifold 52, while the opposite channel opening 74 is in fluid communication with the aperture 82 in the end cover 66 and subsequently with the manifold 54. As will be discussed below, the rotor 46 can rotate in a clockwise direction indicated by the arrow 84. During operation, the low-pressure second fluid 86 passes through the end cover 66 and enters the channel 70, where it comes into contact with the first fluid 88 at the dynamic fluid boundary 90. At this time, the second fluid 86 exits the channel 70, passes through the end cover 64, and expels the first fluid 88 out of the rotating LPC 40. However, due to the short contact time, only minimal mixing exists between the second fluid 86 and the first fluid 88.

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

[0050] In Figure 8, the channel 70 has rotated through an arc of approximately 60 degrees from the position shown in Figure 7. The opening 74 is now in fluid communication with the aperture 80 of the end cover 66, and the opening 72 of the channel 70 is now in fluid communication with the aperture 76 of the end cover 64. At this position, the high-pressure first fluid 88 enters, pressurizing the low-pressure second fluid 86, and causing the second fluid 86 to exit the rotor channel 70 through the aperture 80.

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

[0052] Figure 10 is an exploded assembly view of one embodiment of a rotor 46 having a barrier system 100. As described above, the rotation of the rotor 46 enables pressure transfer between the first and second fluids. To prevent mixing between the first fluid / prime fluid and the second fluid / supercritical fluid within the power generation system 4, the rotary liquid piston compressor 10 includes a barrier system 100. As illustrated, the rotor 46 includes a first rotor section 102 and a second rotor section 104 that are coupled together. By including a rotor 46 having the first and second rotor sections 102, 104, the rotor 46 can accommodate and hold the barrier system 100 inside the rotor 46. As illustrated, the first rotor section 102 includes an end face 106 having an aperture 108 for housing bolts 110. The bolts 110 pass through these apertures 108 and enter the aperture 112 in the second rotor section 104, joining the first and second sections 102 and 104 of the rotor 46. The barrier system 100 is installed between these rotor sections 102 and 104 to securely fasten the barrier system 100 to the rotor 46.

[0053] The barrier system 100 includes a plate 114 to which a plurality of barriers 116 are coupled. These barriers 116 are collapsible diaphragms that prevent contact / mixing between the first and second fluids when pressure is exchanged within the channel 70 of the rotor 46. As discussed below, these barriers 116 expand and contract as pressure is transferred between the first and second fluids. For the purpose of coupling the plate 114 to the rotor 46, the plate 114 may include a plurality of apertures 118 that align with the aperture 108 in the first rotor section 102 and the aperture 112 in the second rotor section 104. These apertures 118 accommodate the bolts 110 when the first rotor section 102 is coupled to the second rotor section 104 and reduce or prevent lateral movement of the plate 114. In some embodiments, the aperture 108 on the first rotor section 102, the aperture 112 on the second rotor section 104, and the aperture 118 on the plate 114 can be positioned on one or more diameters (e.g., inner and outer diameters). In this way, the first rotor section 102 and the second rotor section 104 can compress the plate 114 evenly when joined. In some embodiments, the barriers 116 may not be joined to the plate 114 or may not be supported by the plate 114. Rather, each barrier 116 may be joined to the rotor 46 individually.

[0054] As illustrated, the first rotor section 102 defines a length 120, and the second rotor section 104 defines a length 122. By changing the lengths 120 and 122, the rotor 46 allows for the placement of barrier systems 100 at different locations within the channel 70 along the length of the rotor 46. In this way, the rotary liquid piston compressor 10 can be adapted to various operating conditions. For example, the difference in density and mass flow rate of the two fluids and the rotational speed of the rotor 46 can affect how far the first and second fluids can flow into the channel 70 of the rotor 46 to exchange pressure. Therefore, by changing the lengths 120 and 122 of the first and second rotor sections 102 and 104 of the rotor 46, it becomes possible to place barrier systems 100 at locations that facilitate pressure exchange between the first and second fluids (e.g., midway through the rotor 46).

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

[0056] Figure 11 is a cross-sectional view of one embodiment of a rotor 46 having a barrier system 100. As described above, the barrier system 100 may include plates 114 and barriers 116. These barriers 116 are located inside the channel 70 and block mixing / contact between the first and second fluids while still allowing pressure transfer. To facilitate pressure transfer, the barriers 116 expand and contract. As illustrated in Figure 11, the first barrier 140 of the multiple barriers 116 is in the 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 pushes it out of the rotor 46. Simultaneously, in preparation for pressurization, the second barrier 146 may contract as the second fluid 144 enters the rotor 46. The barrier 116 includes a number of folds 148 (e.g., one, two, three, four, five or more) joined together by ribs 150. It is these elastic folds 148 that allow the barrier 116 to expand in volume as the pressurized first fluid 142 flows into the rotor 46. As will be discussed below, the barrier 116 may be made of one or more materials that provide tensile strength, elongation percentage, and chemical resistance to work with a supercritical fluid (e.g., carbon dioxide).

[0057] Figure 12 is a cross-sectional view of one embodiment of a rotor 46 having a barrier system 100. As illustrated in Figure 12, the first barrier 140 of a plurality of 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 contracts, pressurizing the second fluid 144 and pushing 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 the end 162 (e.g., end portion, end face) and plate 114 of the barrier 116. During operation, the spring 160 stretches as the pressure within the barrier 116 increases, causing the barrier 116 to expand axially 164. Since the spring 160 absorbs force as the barrier 116 expands, the spring 160 can prevent or reduce over-expansion of the barrier 116. Similarly, the spring 160 can also extend the life of the barrier 116, which repeatedly expands and contracts during the operation of the power generation system 4. The spring can also provide a more controlled expansion rate for the barrier 116.

[0058] In some embodiments, the spring 160 can be coupled to the outer surface 168 of the barrier 116 and / or installed outside the barrier 116. In other embodiments, the spring 160 can be coupled to the inner surface 170 and / or installed inside the barrier 116 (i.e., inside the membrane of the barrier 116). In yet another embodiment, the barrier system 100 can include the spring 160 both outside and inside the barrier 116. The spring 160 can also be coupled to the rotor 46 instead of 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] Figure 13 is a cross-sectional view of one embodiment of a rotor 46 having a barrier system 100. In Figure 13, the barrier system 100 includes a planar barrier 190. As illustrated, the planar barrier 190 extends across the channel 70 (for example, in a direction substantially perpendicular to the longitudinal axis of the channel 70) rather than axially into the channel 70, as 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 Figure 13, the first planar barrier 192 of the multiple planar barriers 190 is in the 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 comes into contact with the second fluid 144, pressurizing it and pushing it out of the rotor 46. In preparation for pressurization, the second planar barrier 194 may also be simultaneously contracted as the second fluid 144 enters the rotor 46. The barrier 116 includes a number of connecting folds 196 (e.g., one, two, three, four, five or more). It is these elastic folds 148 that expand as the pressurized first fluid 142 flows into the rotor 46 and contract when the pressure is released.

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

[0061] Figure 15 is a cross-sectional view of one embodiment of the barrier along line 14-14 in Figure 11. As described above, the barriers 116, 190 may be made of one or more materials that provide tensile strength, elongation percentage, and chemical resistance for functioning with supercritical fluids (e.g., the temperature and pressure of the supercritical fluid). In some embodiments, the barriers 116, 119 may include a number of layers (e.g., one, two, three, four, five or more layers) to combine the properties of different materials. For example, the barriers 116, 119 may include two elastomer layers 210 (e.g., ethylene propylene, silicone, nitrile, neoprene, etc.) overlapping the fabric layer 212. During operation, the elastomer layers 210 can provide chemical resistance and high elongation ratio capability, while the fabric layer 212 increases the tensile strength of the barriers 116, 190. Furthermore, one or more of the layers 210 may include a coating 214. The coating 214 may be a coating that has chemical resistance to react with the first fluid and / or the second fluid. For example, layer 210 may include a coating 214 on the outermost surface 216 that chemically protects layer 210 from the supercritical fluid.

[0062] Figure 16 is a cross-sectional view of one embodiment of a rotary liquid piston compressor 10 (e.g., rotary LPC) having a cooling system 240 (i.e., a thermal management system). In some embodiments, the cooling system 240 may include a heat exchanger made of microchannels surrounding the rotary liquid piston compressor. As described above in the description of Figure 1, the fluid changes phase as the temperature and pressure change. At pressures and temperatures above the critical point, 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 the supercritical fluid (e.g., liquid-like density and gaseous viscosity). By controlling the temperature inside the rotary liquid piston compressor 10 having the cooling system 240, the cooling system 240 can prevent the phase change from supercritical fluid to gas phase inside the rotary liquid piston compressor 802. Furthermore, the cooling system 240 can also facilitate the removal of energy as heat is generated during the compression of the supercritical fluid, enabling substantially isothermal compression, which is a thermodynamically more efficient mode of compression. As described above, the cooling system 240 may include microchannels that provide a high surface area per unit volume to facilitate the 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 enclose the housing 244. These conduits 246 may be microconduits having a diameter of 0.05 mm to 0.5 mm. By including microconduits, the cooling system 240 can increase the cooling surface area to control the temperature of the supercritical fluid in the rotary liquid piston compressor 10. The conduits 246 may be arranged in a plurality of transverse rows (e.g., 1, 2, 3, 4, 5 or more) and / or a plurality of vertices (e.g., 1, 2, 3, 4, 5 or more). Each conduit 246 may be fluid-coupled to all other conduits 246, or the cooling system 240 may be coupled to a subset of the conduits 246. For example, all conduits 246 in one 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 fluid-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 enclosed by a housing or covering 247. The housing or covering 247 may be made of a material that blocks and withstands 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 memory 252. For example, the processor 250 may be a microprocessor that runs software for controlling the operation of the actuator 98. The processor 250 may include a number of 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, the processor 250 may include one or more reduced instruction set (RISC) processors.

[0064] Memory 252 may include volatile memory, such as random access memory (RAM), and / or non-volatile memory, such as read-only memory (ROM). Memory 252 can store various types of information and can be used for various purposes. For example, memory 252 can store processor-executable instructions for the processor 250 to execute, such as firmware or software. Memory may include ROM, flash memory, hard drives, or any other optical, magnetic, or solid-state storage medium or a combination thereof. Memory can 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 the cooling fluid source 256 (e.g., a cooling system, an air conditioning system).

[0066] Figure 17 is a cross-sectional view of one embodiment of a rotary liquid piston compressor 802 (RLPC) having a heating system 280 (i.e., a 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 and / or dry ice formation of the fluid 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 enclose 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, cables (single or double) 284 (e.g., coils) carry an electric current that generates heat due to the electrical resistance of the cables 284. The conduits 246 may similarly be surrounded by a housing or covering 286. The housing or covering 286 may be made of a material that blocks and withstands heat transfer, such as polystyrene, fiberglass wool, or various types of foam.

[0068] The flow rate of the heated fluid or current through the conduit or cable 284 is controlled by the controller 248. During operation, the controller 248 may 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 sensors 254 may be installed in direct contact with the supercritical fluid (e.g., inside the cavity housing the supercritical fluid). In some embodiments, the sensors 254 may be installed within the housing 244, sleeve 44, and end covers 64, 66. As the material around the sensors 254 responds to changes in the temperature and / or pressure of the supercritical fluid, the sensors 254 detect this change and transmit it to the controller 248. The controller 248 then correlates this to the actual temperature and / or pressure of the supercritical fluid. Using the feedback from the sensors 254, the controller 248 controls the flow rate of the heated fluid from the heated 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 can 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 examples of supermarket system architectures 300 and 302 that use a rotary pressure exchanger-based supercritical carbon dioxide refrigeration system rather than conventional Joule-Thomson expansion valve-based cooling. In the first architecture 300 (Figure 18), a two-phase low-pressure outflow (e.g., a gaseous / liquid mixture of carbon dioxide) from a rotary pressure exchanger 304 (via a low-pressure outlet 305) proceeds through a flash tank 306 that separates the gas and liquid phases. The liquid carbon dioxide phase is transported to low-temperature (e.g., approximately -20°C) and medium-temperature (e.g., approximately -4°C) heat loads / evaporators 308 and 310 (e.g., the freezer and refrigerator sections of a supermarket, respectively), where the liquid carbon dioxide phase absorbs heat and becomes superheated. Because this is a purely liquid phase rather than a gas / liquid two-phase system, it has a greater heat absorption (i.e., cooling) capacity. Flow control valves 312 and 314 can adjust the flow rate of liquid carbon dioxide to the respective heat loads 308 and 310 (for example, in response to a control signal from a controller). The superheated carbon dioxide vapor from the freezer section 308 then proceeds to the cryogenic compressor 316, where it is reintegrated at the same pressure with the superheated carbon dioxide vapor from the refrigerator section 310 and the separated superheated gaseous carbon dioxide 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 superheated gaseous carbon dioxide flowing from the flash tank 306 (for example, in response to a control signal from a controller). This reintegrated superheated gaseous carbon dioxide then enters the rotary pressure exchanger 304 at the 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 rotating pressure exchanger 304 (via the high-pressure outlet 322) and proceeds to the heat exchanger 324, which is at the highest pressure, where it releases heat into the environment and cools. In some embodiments, the heat exchanger 324 is a gas condenser used for subcritical carbon dioxide.From the gas cooler 324, supercritical carbon dioxide flows to the high-pressure inlet 326 of the rotary pressure exchanger 304. By using a small compressor 328 (e.g., a low-DP circulating 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, it is possible to provide the small pressure rise required to overcome the fluid resistance in the system and the small differential pressure in the rotary pressure exchanger 304.

[0070] The heat exchanger 324 is located along the high-pressure branch to circulate supercritical or subcritical carbon dioxide at high pressure. The cryogenic evaporator 308 and cryogenic compressor 316 are located along the low-pressure branch to circulate carbon dioxide at low pressure (i.e., lower than the pressure in the high-pressure branch) in liquid, gaseous, or vapor state, or in a two-phase mixture of liquid and vapor. The medium-temperature evaporator 310 and valve 314 are located along the intermediate-pressure branch to circulate the refrigerant at an intermediate pressure between the respective pressures of the refrigerant 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 flowing directly out of the flash tank 306 and into the inlet 320 of the rotary pressure exchanger 304 is at the intermediate pressure. Thus, the rotary pressure exchanger 304 is fluid-coupled to the intermediate-pressure branch and the high-pressure branch. The rotary pressure exchanger 304 receives high-pressure refrigerant from the high-pressure branch and intermediate-pressure refrigerant from the intermediate-pressure branch in the form of vapor, liquid, or a two-phase mixture of liquid and vapor, and exchanges pressure between the high-pressure refrigerant and the intermediate-pressure refrigerant. A first outflow of refrigerant at high pressure in a supercritical or subcritical state exits the rotary pressure exchanger, and a second outflow of refrigerant at intermediate pressure in the form of liquid or a two-phase mixture of liquid and vapor exits the rotary pressure exchanger.

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

[0072] The heat exchanger 324 is located along the high-pressure branch to circulate supercritical or subcritical carbon dioxide at high pressure. The cryogenic evaporator 308 and cryogenic compressor 316 are located along the low-pressure branch to circulate carbon dioxide at low pressure (i.e., a pressure lower than the pressure in the high-pressure branch) in the form of liquid, gaseous, or vapor, or as a two-phase mixture of liquid and vapor. The medium-temperature evaporator 310 and valve 314 are located along the first intermediate-pressure branch to circulate refrigerant that is at a first intermediate pressure between the respective pressures of the refrigerant in the low-pressure branch and the second intermediate-pressure branch. The second intermediate-pressure branch is located 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 that flows directly out of the flash tank 306 and into 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. Therefore, the rotary pressure exchanger 304 is fluid-coupled to the second intermediate pressure branch and the high-pressure branch. The rotary pressure exchanger 304 receives high-pressure refrigerant from the high-pressure branch and second intermediate-pressure refrigerant from the second intermediate-pressure branch in the form of vapor, liquid, or a two-phase mixture of liquid and vapor, and exchanges pressure between the high-pressure refrigerant and the second intermediate-pressure refrigerant. Exiting the rotary pressure exchanger are a first exit flow of refrigerant at high pressure in a supercritical or subcritical state, and a second exit flow of refrigerant at second intermediate pressure in the form of liquid or a two-phase mixture of liquid and vapor.

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

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

[0075] In response to the detected flow rate, the controller 584 controls the valve 582 to block and / or reduce the transfer of superheated gaseous carbon dioxide 574 into the working fluid loop 578. For example, if the controller 584 detects a decrease in the flow rate by the sensor 588, the controller 584 can correlate the decreased flow rate with how deeply the working fluid has entered the rotary liquid piston compressor 572 in direction 590. Thus, the controller 584 can determine the associated flow rate of superheated gaseous carbon dioxide 574 into the rotary liquid piston compressor 572 so as to allow the working fluid 580 to exit the rotary liquid piston compressor 572 in direction 592 without allowing the superheated gaseous carbon dioxide 574 to exit the rotary liquid piston compressor 572 in direction 592. In other words, the controller 584 controls the valve 582 to block the flow of superheated gaseous carbon dioxide 574 into the working fluid loop 578 by ensuring that the flow rate of the working fluid 580 into the rotary liquid piston compressor 572 is greater than the flow rate of superheated gaseous carbon dioxide 574.

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

[0077] Figure 21 is a schematic diagram of one embodiment of a control system 620 that controls the movement of a fluid (e.g., supercritical carbon dioxide, superheated gaseous carbon dioxide) in a rotary liquid piston compressor 622. As described above, energy can be exchanged between two fluids using a rotary liquid piston compressor or pump. For example, the rotary liquid piston compressor 622 can be used to exchange energy between two fluids in the refrigeration system described above. With the aim of reducing the transfer of superheated gaseous carbon dioxide 624 or a two-phase gas / liquid carbon dioxide mixture 625 in the fluid loop 626 and / or preventing it from entering the working fluid loop 628 through which the working fluid 630 (e.g., superheated carbon dioxide) circulates, the control system 620 can control the distance that carbon dioxide travels axially within the rotor channel in the rotary liquid piston compressor 622 in response to the flow rate of the working fluid 630 and the flow rate of superheated gaseous carbon dioxide 624. The control system 620 controls the movement of the drive fluid by decelerating or accelerating the rotational speed of the rotor of the rotary liquid piston compressor 622. In other words, by controlling the rotational speed, the control system 620 can prevent and / or limit the superheated gaseous carbon dioxide 624 from flowing completely through the rotating liquid piston compressor 622 (i.e., completely through the channel 70 shown in Figure 5) and into the working fluid loop 628.

[0078] To reduce the mixing of the working fluid 630 with 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 as shown in Figure 5) and, consequently, the axial length to which the superheated gaseous carbon dioxide 624 can flow into the rotor channels. The faster the rotor rotates, the shorter the time the superheated gaseous carbon dioxide and working fluid must flow into the rotor channels, thus reducing 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 the superheated gaseous carbon dioxide and working fluid must flow into the rotor channels, thus increasing 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 that detect the flow rates of the superheated gaseous carbon dioxide 624 and the working fluid 630, respectively, and emit signals indicating the flow rates. The controller 638 receives and processes signals for detecting 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 a command to the variable frequency drive to control the speed of the motor 632 in order to block and / or reduce the transfer of 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 with how far the working fluid has moved in direction 640 into the channel of the rotary liquid piston compressor 622. Therefore, the controller 638 can determine the associated speed of the motor 632 that expels the working fluid 630 from the rotary liquid piston compressor 622 in direction 642 without expelling the superheated gaseous carbon dioxide 624 from the rotary liquid piston compressor 622 in direction 642.

[0080] In response to the lower instantaneous flow rate of the working fluid compared to the superheated gaseous carbon dioxide, the controller 638 controls the motor 632 through a variable frequency drive to increase the rotational speed of the rotary liquid piston compressor 622 (i.e., increase the revolutions per minute) to shorten the axial length over which 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 excessively high compared to the drive fluid, the controller 638 reduces the rotational speed of the rotary liquid piston compressor 622 to increase the axial distance over which the superheated gaseous carbon dioxide 624 travels within the channels of the rotary liquid piston compressor 622 to expel the working fluid 630 from the rotary liquid piston compressor 622.

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

[0082] As noted above, since almost all of the compression takes place inside the rotary pressure exchanger, in some embodiments it is possible to completely or partially eliminate the main compressor (e.g., bulk flow compressor). For example, the compressor can be replaced by a gas blower with a very low differential pressure or a circulation pump that consumes very little work (because the enthalpy change across it is very small). Figure 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., having a low-flow, high-DP leakage pump and a low-DP, high-flow circulation pump instead of a bulk flow compressor). In general, the refrigeration system 900 is similar to the refrigeration system 800 in Figure 2.

[0083] As depicted, the 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., for cooling or heat load), a low-pressure, high-flow, low-DP multi-phase circulation pump 911, and the low-pressure side of a rotary pressure exchanger 902. The rotary pressure exchanger 902 fluidly couples the high-pressure and low-pressure loops 904 and 906. Furthermore, the multiphase flow leak pump 913, operating at a low flow rate but high DP, extracts any leaks from the low-pressure pressure exchanger 902 at its low-pressure outlet 920 and returns them to the high-pressure loop 904 just upstream of the high-pressure inlet 914 of the pressure exchanger 902. The multiphase flow pump 909 in the high-pressure loop 904 ensures the maintenance of the required flow rate within the high-pressure loop 904 by overcoming small pressure losses within the loop 904. Because there is not much differential pressure across pump 909, this pump consumes very little energy. The flow rate entering this multiphase flow pump 909 is from the outlet 936 of the gas cooler / condenser 908 and can 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 pump 909, the flow rate exiting pump 909 is considered to be the same as the inflow rate entering the high-pressure inlet 914 of the pressure exchanger 902 at that time. The flow rate from the low-pressure outlet 920 of the pressure transducer 902 is thought to be in a two-phase liquid-vapor state or a pure liquid state.

[0084] The multiphase flow pump 911 in the low-pressure loop 906 circulates this bulk low-pressure flow rate of the refrigerant through the evaporator 910 and sends it to the low-pressure inlet 918 of the pressure transducer 902. The multiphase flow pump 911 also has very little differential pressure across it (i.e., just 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 flow pump 911 circulates the flow rate through the evaporator 910, where it gains heat and converts itself into a pure vapor state or a two-phase liquid-vapor mixture with a higher vapor content. This high vapor-content flow rate then enters the low-pressure inlet 918 of the pressure transducer 902 and is pressurized to high pressure. This, in turn, raises the temperature of the fluid as well, according to the standard laws of thermodynamics. This higher-pressure, hotter fluid then exits through the high-pressure outlet 922 of the pressure transducer 902. The fluid exiting from the high-pressure outlet 922 may be in a supercritical state, or depending on how the system is optimized, it may exist in a subcritical state or as a mixture of liquid and vapor with a high vapor content. This high-pressure, high-temperature refrigerant then enters the gas cooler / condenser 908 of the high-pressure loop 904, where it releases heat into the surrounding environment. By releasing heat, the refrigerant cools (if in a supercritical state) or changes phase to a liquid state. The multiphase flow pump 909 in the high-pressure loop 904 then receives this liquid refrigerant and circulates it through the high-pressure loop 904 as described earlier.

[0085] If there is no internal leak within the pressure transducer 902, the high-pressure loop 904 will remain at a constant high pressure, and the low-pressure loop 906 will remain at a constant low pressure. However, if there is an internal leak within the pressure transducer 902 from the high-pressure side to the low-pressure side, a net flow transfer from the high-pressure loop 904 to the low-pressure loop 906 is considered to exist. To account for this transfer and pump this leaked flow back into the high-pressure loop 904, a third multiphase flow pump 913, which is a high differential pressure, low flow rate leak pump, is used. Pump 913 takes up any excess flow leaking into the low-pressure loop 906 at low pressure and pumps it back into the high-pressure loop 904 to maintain mass equilibrium and pressure in each loop 904 and 906. Within the low-pressure loop 906, a three-way valve 915 is positioned between the low-pressure outlet 920 of the pressure transducer 902 and the inlet of the low-pressure multiphase flow pump 911. Valve 915 divides the flow rate, allowing only the excess flow rate coming from the low-pressure outlet 920 of the pressure transducer 902 to be directed to the high-DP multiphase flow pump 913. Pump 913 also allows pumping any additional flow rate coming from the low-pressure outlet 920, due to the compressibility of the refrigerant and the density difference between the four flows entering and leaving the pressure transducer 902. Pump 913 also helps maintain the pressure in the low-pressure loop 906 at a constant low pressure and the pressure in the high-pressure loop 904 at a constant high pressure. Between the outlet of the high-pressure multiphase flow pump 909 and the high-pressure inlet 94 of the pressure transducer 902, another three-way valve 917 is located within the high-pressure loop 904. Valve 917 allows the leak / excess flow rate originating from the high-DP multiphase flow pump 913 to be combined with the high-pressure bulk flow rate coming from the high-pressure multiphase flow pump 909 and then directed into the high-pressure inlet 914 of the pressure transducer 902. Although the differential pressure across the multiphase flow pump 913 is high, the flow rate it must pump is very small (for example, about 1-10 percent of the bulk flow rate that goes through either of the other two pumps 909 or 911). Therefore, the energy consumption of pump 913 is also relatively low.When the energy consumption of all three multiphase flow pumps 909, 911, and 913 is added together, it still appears to be far 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). This is the main advantage of this configuration.

[0086] Figure 22B demonstrates another embodiment of the refrigeration system 923 without a bulk flow compressor. It is similar to system 900 shown in Figure 22A, except that any excess flow exiting from the low-pressure outlet 920 of the pressure transducer 902 (due to an internal leak in the pressure transducer 902 or, as previously described, due to the compressibility and density differences of the four flows entering and exiting the pressure transducer 902) is pumped through the evaporator 910 along with the valved low-pressure flow and converted to steam before being compressed and returned into the high-pressure loop 904. Thus, the high-DP, low-flow multiphase flow leak pump 913 in Figure 22A is replaced by a high-DP, low-flow leak compressor 925 as shown in Figure 22B. The leak compressor 925 compresses the excess flow from a low-pressure steam state to a high-pressure steam state, or to a supercritical state, before injecting it into the high-pressure loop 904. The location of this reinjection of the surplus flow rate is similarly different from that shown in Figure 22A. The vaporized or supercritical refrigerant 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 outlet pressure of the leak compressor). As shown in Figure 22B, a three-way valve 927 is positioned downstream of the evaporator 910 to allow the surplus flow rate from the bulk flow rate in the low-pressure loop 906 to be split before being sent through the leak compressor 925. Similarly, a three-way valve 929 is positioned downstream of the pressure transducer 902 to allow the recombination of the high-pressure leak flow rate exiting the leak compressor 925 and the high-pressure bulk flow rate exiting the pressure transducer 902. This combined high-pressure flow rate then proceeds to the gas cooler / condenser 908, as previously described. The advantage of this configuration compared to the one in Figure 22A is that it provides additional heat absorption capacity to the cycle due to the additional flow rate passing through the evaporator 910 (excess flow rate 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 system 900 shown in Figure 22A, because the energy consumed by the leak compressor 925 is slightly higher than the energy consumed by the multiphase flow leak pump 913.This is because, unlike the multiphase flow circulation pump 913 where the refrigerant is pumped in a partially or completely liquid state, in the leak compressor 925 it is compressed to high pressure in a completely vapor state.

[0087] The thermodynamic processes occurring within the refrigeration system 923 are described in more detail with reference to Figures 23 and 24. Figures 23 and 24 illustrate the thermodynamic processes occurring in four main components of the refrigeration system 900, illustrating the temperature-entropy (TS) diagram 926 and the pressure-enthalpy (PH) diagram 928, respectively. 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, consequently, the temperature of the refrigerant working fluid (e.g., carbon dioxide) to a temperature higher than the ambient temperature, where it can release heat into the hotter external environment. This occurs inside the gas cooler 908. In the gas cooler 908 of the supercritical carbon dioxide system, since 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, as the carbon dioxide releases heat into the environment, the temperature drops. As illustrated in Figures 23 and 24, the supercritical carbon dioxide at the gas cooler outlet 936 enters the rotary pressure exchanger 902 at the high-pressure inlet port 914, undergoes isentropic or near-isentropic expansion (approximately 85 percent isentropic efficiency), 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, constant enthalpy process). Overall, diagrams 926 and 928 illustrate the benefits of cycle efficiency resulting from increased cooling capacity and reduced compressor workload. Since the expansion inside the rotating 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 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] Figure 25 is a schematic diagram of a refrigeration system 931 that uses a low-DP circulating compressor instead of a circulating pump. The circulating compressor overcomes the minimum pressure loss in system 931 by maintaining fluid flow rate throughout system 900. The difference between this system and systems 900 and 923 shown in Figures 22A and 22B is that the circulation of bulk flow in the low-pressure loop 906 and high-pressure loop 904 is achieved using a low-DP circulating compressor instead of a low-DP multiphase flow circulating pump. Similarly, the locations of these circulating compressors are also different. For example, circulating compressor 941 (compressor 1) in the low-pressure loop 906 is located downstream of the evaporator 910, which circulates the refrigerant in a vapor state. Similarly, circulating compressor 944 (compressor 2) in the high-pressure loop 904 is located downstream of the high-pressure outlet 922 of the pressure transducer 902, which circulates the refrigerant in a supercritical or high-pressure vapor state. Compressor 3 is similar to the high-DP, low-flow leak compressor 925 described in relation to Figure 22B, where compressor 925 takes the excess flow (e.g., leak flow from pressure transducer 902) entering the low-pressure loop 906 from pressure transducer 902 in a vapor state and compresses it back into the high-pressure loop 904 either as a high-pressure vapor state or in a supercritical state. This excess flow is then combined with the high-pressure bulk flow coming from compressor 944 before proceeding to the gas cooler / condenser 908. A low-DP circulating compressor 941 positioned along the second fluid loop 906 (e.g., the low-pressure fluid loop) maintains the fluid flow along loop 906 (e.g., between the rotating pressure exchanger 902 and the gas cooler 908). Furthermore, a low-DP circulating compressor 944 positioned along the first fluid loop 904 (e.g., the high-pressure fluid loop) maintains the fluid flow along loop 904 (e.g., between the evaporator 910 and the rotating pressure exchanger 902). In some embodiments, the refrigeration system 931 may include only compressors 925 and 941. In some embodiments, the refrigeration system 900 may include only compressors 944 and 941. In some embodiments, compressors 941 and 944 each have a significantly lower differential pressure across them than the leak compressor 925, as will be noted in more detail below.

[0089] In some embodiments, a three-way valve is located at the junction between the flows exiting from compressors 925 and 944 (for example, near circle 2 in Figure 25). This three-way valve is located between the high-pressure, high-flow, and low-DP circulating compressor 944 and the gas cooler or condenser 908 within the high-pressure branch 904, where, during the operation of the refrigeration system 931, the first flow from the high-DP, low-flow leak compressor 925 is combined with the bulk flow exiting the high-pressure, high-flow, and low-DP circulating compressor 944 before proceeding to the inlet 934 of the gas cooler or condenser 908. The high-pressure, high-flow, and low-DP circulating 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 of the evaporator 910 branching toward the compressors 925, 941 (for example, near 1 in the circle in Figure 25). This three-way valve is located between the evaporator 910 and the rotary pressure exchanger 902 in the low-pressure branch 906, where, during the operation of the refrigeration system 931, a portion of the flow rate leaving the evaporator 910 is diverted through the three-way valve to the inlet of the high-DP, low-flow leak compressor 925, while the remaining portion of the flow rate proceeds to the low-pressure inlet 918 of the rotary pressure exchanger 902. A low-pressure, high-flow, and low-DP circulating 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 113.56 liters (30 gallons) per minute and a differential pressure of approximately 10,342 kPa (1,500 psi). Assuming these operating conditions, the bulk flow compressor is estimated to require approximately 45,000 (i.e., 30 × 1,500 psi) units of force (i.e., work done or energy consumed). In the refrigeration system 900 described above, the low-DP circulating compressors 941 and 944 (assuming each operates at a flow rate of approximately 113.56 liters (30 gallons) per minute and a differential pressure of approximately 68.9 kPa (10 psi)) are estimated to require approximately 300 (i.e., 30 × 10) units of force, respectively. The leak compressor 925 (assuming it operates at a flow rate of approximately 5.68 liters (1.5 gallons) and a differential pressure of approximately 10,342 kPa (1,500 psi)) is estimated to require approximately 2,250 (i.e., 1.5 × 1,500) units of force. Therefore, the compressors 925, 941, and 944 in the refrigeration system 931 are estimated to require approximately 2,850 units of force. Thus, the compressors 925, 941, and 944 are estimated to reduce energy consumption by at least one-tenth (and even up to one-fifteenth) compared to a bulk-flow compressor-based system.

[0092] In some embodiments, a refrigeration system 931 (having one or more of a leak compressor 925 and low-DP circulating compressors 941, 944) may be used in the supermarket architecture described above in Figures 18 and 19.

[0093] Figures 26 and 27 show two examples of supermarket system architectures 950, 952 that use 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. Furthermore, although architectures 950, 952 are discussed in relation to the use of a gas cooler as a heat exchanger 324 for use with supercritical refrigerants (e.g., carbon dioxide), it is also possible to use these architectures 950, 952 with a condenser as a heat exchanger 324 for use with subcritical refrigerants (e.g., carbon dioxide). In the first architecture 950 (Figure 26), a two-phase low-pressure outflow from a rotary pressure exchanger 304 (via a low-pressure outlet 305) (e.g., a gas / liquid mixture of carbon dioxide at a first intermediate pressure such as 370 psi) proceeds through a flash tank 306 that separates the gas phase and the liquid phase (both exiting the flash tank at e.g., 370 psi). The liquid carbon dioxide phase is transported to low-temperature (e.g., approximately -20°C) and medium-temperature (e.g., approximately -4°C) heat load / evaporators 308 and 310 (e.g., the freezer and refrigerator sections of a supermarket, respectively), where the liquid carbon dioxide phase absorbs heat and becomes superheated. Because this is purely a liquid phase rather than a two-phase gas / liquid phase, it has a greater heat absorption (i.e., cooling) capacity. The liquid carbon dioxide phase enters the medium-temperature evaporator 310 at, for example, 370 psi, while the liquid carbon dioxide phase flows through the flow control valve 312 and then enters the low-temperature evaporator 308 at, for example, 180 psi. The flow control valve 312 can regulate the flow rate of liquid carbon dioxide to the evaporator 308 (e.g., in response to a control signal from a controller). The superheated carbon dioxide vapor from the freezer section 308 (at a low pressure of 180 psi) then proceeds to the cryogenic compressor 316 (where it exits at, for example, 370 psi), after which it recombines with the superheated carbon dioxide vapor from the refrigerator section 310 (at, for example, 370 psi) and the separated superheated gaseous carbon dioxide separated from the gas / liquid mixture in the flash tank 306, which is at the same pressure.A control valve 318 (e.g., a flash gas control valve) can regulate the flow rate of superheated gaseous carbon dioxide flowing from the flash tank 306 (e.g., in response to a control signal from a controller). This reintegrated superheated gaseous carbon dioxide then enters the rotary pressure exchanger 304 at the low-pressure inlet port 320 and is compressed to a second intermediate pressure (e.g., 500 psi). The superheated gaseous carbon dioxide exits the rotary pressure exchanger 304 (via the high-pressure outlet 322) and proceeds to the medium-temperature compressor 330, where the superheated gaseous carbon dioxide is compressed to the maximum pressure in the system (e.g., 1,300 psi) depending on the system requirements and converted to supercritical carbon dioxide. The supercritical carbon dioxide then proceeds at the maximum pressure to the heat exchanger 324 (e.g., a gas cooler), where it releases heat to the environment and is cooled. In some embodiments, the heat exchanger 324 is a gas condenser used for subcritical carbon dioxide. From the gas cooler 324, supercritical carbon dioxide (e.g., at 1,300 psi) flows through the 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 is a slight variation from architecture 950 in Figure 26. More specifically, as depicted in Figure 27, the carbon dioxide gas / liquid mixture (at a second intermediate pressure, e.g., 500 psi) flows into the flash tank 306 for separation into pure carbon dioxide gas or vapor and liquid. The carbon dioxide gas from the flash tank 306 flows into the high-pressure inlet 326 of the rotary pressure exchanger 304, while the carbon dioxide liquid from the flash tank flows into the low-pressure chambers 308 and 310. The two-phase gas-liquid CO2 mixture exiting the low-pressure inlet 305 of the pressure transducer 304 exits at the same pressure as the medium-temperature evaporator 310 and, after combining with the fluid stream exiting the medium-temperature evaporator 310 and the cryogenic compressor 316, enters the low-pressure inlet 320 of the pressure transducer 304. Similarly, the flow control valve 314 is located upstream of the medium-temperature evaporator 310.

[0095] Although the present invention may be subject to various modifications and variations, specific embodiments are shown in the drawings as examples and are described herein. However, it should be understood that the present invention is not intended to be limited to any particular form disclosed. Rather, the present invention should encompass all modifications, equivalents, and variations that fall within the spirit and scope of the invention as defined by the following appended claims. The following are embodiments of the present invention. [Aspect 1] In a refrigeration system, A high-pressure branch for circulating the refrigerant at high pressure through the inside; A gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure branch is configured to release heat from the high-pressure refrigerant to the surrounding environment via the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state; A second low-pressure branch for circulating the refrigerant at low pressure through the middle; An evaporator arranged along the low-pressure branch, wherein the low-pressure branch is configured to absorb heat from the surrounding environment into the low-pressure refrigerant via the evaporator, and the low-pressure refrigerant is 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 low pressure to high pressure; A rotary pressure exchanger fluidly coupled to the low-pressure branch and the high-pressure branch, configured to receive high-pressure refrigerant from the high-pressure branch, receive low-pressure refrigerant from the low-pressure branch, and exchange pressure between the high-pressure refrigerant and the low-pressure refrigerant, wherein a first outflow from the rotary pressure exchanger contains high-pressure refrigerant in a supercritical or subcritical state, and a second outflow from the rotary pressure exchanger contains low-pressure refrigerant in a liquid state or a two-phase mixture of liquid and vapor; A refrigeration system including a refrigeration system. [Aspect 2] The refrigeration system according to embodiment 1, wherein the refrigerant contains carbon dioxide. [Aspect 3] The refrigeration system according to embodiment 1 or 2, wherein the rotating pressure exchanger is configured to enable the compression of the received low-pressure refrigerant, which is in a vapor state or a two-phase mixture of liquid and vapor, into a high-pressure refrigerant in a supercritical or subcritical state, and to enable the expansion of the received high-pressure refrigerant in a supercritical or subcritical state into a two-phase mixture of liquid and vapor or a low-pressure refrigerant in a liquid state. [Aspect 4] The evaporator is located downstream of the rotary pressure exchanger and is configured to receive the low-pressure refrigerant, which is 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. The refrigeration system according to embodiment 3. [Aspect 5] A refrigeration system according to any one of embodiments 1 to 4, comprising the compressor fluidly coupled to the low-pressure and high-pressure branches. [Aspect 6] The refrigeration system according to embodiment 5, wherein the evaporator is configured to supply a first portion of the low-pressure refrigerant in vapor state to the rotary pressure exchanger and a second portion of the low-pressure refrigerant in vapor state to the compressor, and the first and second portions of the low-pressure refrigerant in vapor state include superheated steam. [Aspect 7] A refrigeration system according to any one of embodiments 1 to 6, wherein the rotating pressure exchanger is configured to expand the high-pressure refrigerant in a supercritical state into a low-pressure refrigerant which is a two-phase mixture of liquid and vapor via isentropic expansion or quasi-isentropic expansion. [Aspect 8] A refrigeration system according to any one embodiment 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 workload of the compressor. [Aspect 9] In a refrigeration system, A high-pressure branch for circulating the refrigerant at high pressure through the inside; A gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure branch is configured to release heat from the high-pressure refrigerant to the surrounding environment via the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state; A low-pressure branch for circulating the refrigerant at low pressure through the inside; A first evaporator is arranged along the low-pressure branch, the first evaporator is configured to operate at a first temperature, the low-pressure branch is configured to absorb heat from the ambient environment into the low-pressure refrigerant via the evaporator, the low-pressure refrigerant 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 through a first intermediate pressure; A second evaporator, positioned along the first intermediate pressure branch, and configured to operate at a second temperature higher than the first temperature; A second intermediate pressure branch for circulating the refrigerant through a second intermediate pressure, wherein the first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the pressures of 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 the saturation pressure in the second evaporator, and the second intermediate pressure of the refrigerant in the second intermediate pressure branch is between the pressures of the high-pressure branch and the first intermediate pressure branch; A flash tank configured to operate at the second intermediate pressure and separate the refrigerant, which is a two-phase mixture of liquid and vapor, 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, wherein the rotary pressure exchanger is configured to receive the high-pressure refrigerant from the high-pressure branch, receive the second intermediate-pressure refrigerant from the second intermediate-pressure branch in the form of vapor, liquid, or a two-phase mixture of liquid and vapor, and exchange pressure between the high-pressure refrigerant and the second intermediate-pressure refrigerant, wherein a first outflow from the rotary pressure exchanger contains the high-pressure refrigerant in the supercritical or subcritical state, and a second outflow from the rotary pressure exchanger contains the second intermediate-pressure refrigerant in the form of liquid or a two-phase mixture of liquid and vapor; A refrigeration system including a refrigeration system. [Aspect 10] The refrigeration system according to embodiment 9, comprising a first compressor positioned downstream of the flash tank and the first evaporator, wherein the first compressor operates at a first temperature and is configured to receive the refrigerant from the first evaporator, which is either in vapor state or a two-phase mixture of liquid and vapor, and to pressurize the refrigerant to a first intermediate pressure. [Aspect 11] A refrigeration system according to embodiment 10, comprising a second compressor positioned downstream of the first compressor and the second evaporator, wherein the second compressor operates at a second temperature, and the second compressor is configured to receive a refrigerant from both the first compressor and the second evaporator, which is either in vapor state or a two-phase mixture of liquid and vapor, and to pressurize the refrigerant to the high pressure. [Aspect 12] A refrigeration system according to any one embodiment of embodiments 9 to 11, comprising a first valve configured to regulate the flow rate of separated liquid refrigerant from the flash tank so that it flows to the first evaporator after the separated liquid refrigerant reaches the low pressure. [Aspect 13] The refrigeration system according to embodiment 12, further comprising a second valve configured to regulate the flow rate of the separated liquid refrigerant from the flash tank to the second evaporator after the separated liquid refrigerant has reached the first intermediate pressure. [Aspect 14] The refrigeration system according to embodiment 13, further comprising a third valve configured to regulate the flow rate of separated vapor refrigerant from the flash tank at the second intermediate pressure to the inlet of the rotating pressure exchanger. [Aspect 15] A refrigeration system according to any one embodiment of embodiments 9 to 14, wherein the refrigerant contains carbon dioxide. [Aspect 16] In a refrigeration system, A high-pressure branch for circulating the refrigerant at high pressure through the inside; A gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure branch is configured to release heat from the high-pressure refrigerant to the surrounding environment via the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state; A low-pressure branch for circulating the refrigerant at low pressure through the inside; A first evaporator, positioned along the low-pressure branch, is configured to operate at a first temperature, and the low-pressure branch is configured to absorb heat from the ambient environment into the low-pressure refrigerant via the evaporator, wherein the low-pressure refrigerant is in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; An intermediate pressure branch for circulating the refrigerant through the middle at an intermediate pressure; A second evaporator arranged along the intermediate pressure branch, configured to operate at a second temperature higher than the first temperature, wherein the 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 the saturation pressure in the second evaporator; A flash tank configured to operate at the aforementioned intermediate pressure and to separate the refrigerant, which is a two-phase mixture of liquid and vapor, into a pure liquid and a pure vapor; A rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high-pressure branch, wherein it receives the high-pressure refrigerant from the high-pressure branch, receives the intermediate-pressure refrigerant from the intermediate-pressure branch in the form of vapor, liquid, or a two-phase mixture of liquid and vapor, and is configured to exchange pressure between the high-pressure refrigerant and the intermediate-pressure refrigerant, wherein a first outflow from the rotary pressure exchanger contains the high-pressure refrigerant in a supercritical or subcritical state, and a second outflow from the rotary pressure exchanger contains the intermediate-pressure refrigerant in the form of liquid or a two-phase mixture of liquid and vapor; A refrigeration system including a refrigeration system. [Aspect 17] A refrigeration system according to embodiment 16, comprising a low differential pressure compressor configured to receive the refrigerant in a supercritical or subcritical state as it exits the rotary pressure exchanger, and to pressurize the refrigerant to the high pressure. [Aspect 18] A refrigeration system according to embodiment 17, comprising a flash tank and a compressor positioned downstream of the first evaporator, wherein the first compressor operates at a first temperature, and the compressor is configured to receive the refrigerant from the first evaporator in the vapor state or as a two-phase mixture of liquid and vapor, and to pressurize the refrigerant to the intermediate pressure for the rotating pressure exchanger. [Aspect 19] A refrigeration system according to embodiment 18, comprising a valve configured to regulate the flow rate of separated vapor refrigerant from the flash tank at the intermediate pressure up to the inlet of the rotary pressure exchanger. [Aspect 20] A refrigeration system according to any one embodiment of embodiments 16 to 19, wherein the refrigerant contains carbon dioxide.

Claims

1. In a refrigeration system, A high-pressure branch for circulating the refrigerant at high pressure through the inside; A gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure branch is configured to release a first heat to a first ambient environment from the high-pressure refrigerant via the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state; A low-pressure branch for circulating the refrigerant at low pressure through the inside; A first evaporator is arranged along the low-pressure branch, the first evaporator is configured to operate at a first temperature, and the low-pressure branch is configured to absorb a second heat from a second ambient environment into the low-pressure refrigerant via the first evaporator, the low-pressure refrigerant 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 through a first intermediate pressure; A second evaporator, positioned along the first intermediate pressure branch, and configured to operate at a second temperature higher than the first temperature; A second intermediate pressure branch for circulating the refrigerant through a second intermediate pressure, wherein the first intermediate pressure of the refrigerant in the first intermediate pressure branch is between the 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 the saturation pressure in the second evaporator, and the second intermediate pressure of the refrigerant in the second intermediate pressure branch is between the pressures of the refrigerant in the high-pressure branch and the refrigerant in the first intermediate pressure branch; A flash tank configured to operate at the second intermediate pressure and separate the refrigerant, which is a two-phase mixture of liquid and vapor, into substantially pure liquid and pure vapor; A rotary pressure exchanger fluidly coupled to the second intermediate pressure branch and the high-pressure branch, wherein the rotary pressure exchanger includes a rotor that forms a plurality of channels, and the rotor is The high-pressure refrigerant is received from the high-pressure branch into one or more channels. From the second intermediate pressure branch, the refrigerant at the second intermediate pressure, which is in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor, is received into at least one of the channels. Furthermore, the rotary pressure exchanger is configured to exchange pressure between the high-pressure refrigerant and the second intermediate-pressure refrigerant via the rotor of the rotary pressure exchanger, wherein the first outflow from the rotary pressure exchanger contains the high-pressure refrigerant in the supercritical or subcritical state, and the second outflow from the rotary pressure exchanger contains the second intermediate-pressure refrigerant in the liquid state or a two-phase mixture of liquid and vapor; A refrigeration system including a refrigeration system.

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

3. The refrigeration system according to claim 2, further comprising a second compressor positioned downstream of the first compressor and the second evaporator, wherein the second compressor operates at a second temperature, and the second compressor is configured to receive a refrigerant, which is either in vapor state or a 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. The refrigeration system according to claim 1, further comprising a first valve configured to regulate a first flow rate of separated liquid refrigerant from the flash tank such that the separated liquid refrigerant flows to the first evaporator after reaching the low pressure.

5. The refrigeration system according to 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 has reached the first intermediate pressure.

6. The refrigeration system according to 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 the inlet of the rotary pressure exchanger.

7. The refrigeration system according to claim 1, wherein the refrigerant includes carbon dioxide.

8. In a refrigeration system, A high-pressure branch for circulating the refrigerant at high pressure through the inside; A gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure branch is configured to release a first heat to a first ambient environment from the high-pressure refrigerant via the gas cooler or condenser, and the high-pressure refrigerant is in a supercritical or subcritical state; A low-pressure branch for circulating the refrigerant at low pressure through the inside; A first evaporator is arranged along the low-pressure branch, the first evaporator is configured to operate at a first temperature, and the low-pressure branch is configured to absorb a second heat from a second ambient environment into the low-pressure refrigerant via the first evaporator, the low-pressure refrigerant being in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; An intermediate pressure branch for circulating the refrigerant through the middle at an intermediate pressure; A second evaporator arranged along the intermediate pressure branch, configured to operate at a second temperature higher than the first temperature, wherein the intermediate pressure of the refrigerant in the intermediate pressure branch is between the pressures of the refrigerant in the high-pressure branch and the refrigerant in the low-pressure branch, and the intermediate pressure of the refrigerant in the intermediate pressure branch is substantially equal to the saturation pressure in the second evaporator; A flash tank configured to operate at the aforementioned intermediate pressure and separate the refrigerant, which is a two-phase mixture of liquid and vapor, into substantially pure liquid and pure vapor; A rotary pressure exchanger fluidly coupled to the intermediate pressure branch and the high-pressure branch, wherein the rotary pressure exchanger is The high-pressure refrigerant is received from the high-pressure branch. From the intermediate pressure branch, the refrigerant at the intermediate pressure is received, which is in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor. A rotating pressure exchanger configured to exchange pressure between the high-pressure refrigerant and the intermediate-pressure refrigerant, wherein the first outflow from the rotating pressure exchanger contains the high-pressure refrigerant in a supercritical or subcritical state, and the second outflow from the rotating pressure exchanger contains the intermediate-pressure refrigerant in a liquid state or a two-phase mixture of the liquid and vapor; A refrigeration system including a refrigeration system.

9. The refrigeration system according to claim 8, further comprising a low differential pressure compressor configured to receive the refrigerant in a supercritical or subcritical state as it exits the rotary pressure exchanger, and to pressurize the refrigerant to the high pressure.

10. The refrigeration system according to claim 9, further comprising a flash tank and a second compressor positioned downstream of the first evaporator, wherein the second compressor is configured to operate at the first temperature, and the second compressor is configured to receive the refrigerant from the first evaporator, which is either in vapor state or a two-phase mixture of liquid and vapor, and to pressurize the refrigerant to the intermediate pressure for the rotating pressure exchanger.

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

12. The refrigeration system according to claim 8, wherein the refrigerant includes carbon dioxide.

13. Circulating the refrigerant at high pressure to the high-pressure branch; Releasing first heat from the high-pressure refrigerant to the first ambient environment via a gas cooler or condenser arranged along the high-pressure branch, wherein the high-pressure refrigerant is in a supercritical or subcritical state; Circulating the refrigerant at low pressure through a low-pressure branch; Absorbing a second heat from a second ambient environment into the low-pressure refrigerant via a first evaporator arranged along the low-pressure branch, wherein the first evaporator is configured to operate at a first temperature, and the low-pressure refrigerant is in a liquid state, a vapor state, or a two-phase mixture of liquid and vapor; The refrigerant is circulated to the intermediate pressure branch at an intermediate pressure; Absorbing a third heat from a third ambient environment into the intermediate-pressure refrigerant via a second evaporator arranged along the intermediate-pressure branch, wherein the second evaporator is configured to operate at a second temperature greater than the first temperature, the intermediate pressure of the refrigerant in the intermediate-pressure branch is between the pressures of the refrigerant in the high-pressure branch and the refrigerant in the low-pressure branch, and the intermediate pressure of the refrigerant in the intermediate-pressure branch is substantially equal to the saturation pressure in the second evaporator, and circulating the refrigerant through the intermediate-pressure branch; The refrigerant, which is a two-phase mixture of liquid and vapor, is separated into substantially pure liquid and pure vapor via a flash tank operating at the aforementioned intermediate pressure; The high-pressure refrigerant is received 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 an intermediate pressure, which is in the vapor state, the liquid state, or a two-phase mixture of the liquid and vapor, via the rotary pressure exchanger; Exchanging pressure between a high-pressure refrigerant and an intermediate-pressure refrigerant via the rotary pressure exchanger, wherein a first outflow from the rotary pressure exchanger contains the high-pressure refrigerant in a supercritical or subcritical state, and a second outflow from the rotary pressure exchanger contains the intermediate-pressure refrigerant in a liquid state or a two-phase mixture of liquid and vapor; A method that includes this.

14. Receiving the refrigerant in a supercritical or subcritical state as it exits the rotating pressure exchanger via a low differential pressure compressor; The method according to claim 13, further comprising pressurizing the refrigerant to the high pressure via the low differential pressure compressor.

15. Receiving the refrigerant from the first evaporator, either in vapor form or as a two-phase mixture of liquid and vapor, via the flash tank and a second compressor positioned downstream of the first evaporator, wherein the second compressor is configured to operate at the first temperature; The method according to claim 14, further comprising pressurizing the refrigerant to the intermediate pressure for the rotary pressure exchanger via the second compressor.

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

17. The method according to claim 13, wherein the refrigerant contains carbon dioxide.

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