US20120234026A1

US20120234026A1 - High efficiency refrigeration system and cycle

Info

Publication number: US20120234026A1
Application number: US13/453,600
Authority: US
Inventors: JongSik Oh; Gridhari L. Agrawal; Charles William Buckley; Samuel T. Browning
Original assignee: R&D Dynamics Corp
Current assignee: R&D Dynamics Corp
Priority date: 2009-06-10
Filing date: 2012-04-23
Publication date: 2012-09-20

Abstract

A high efficiency air conditioning and refrigeration system and cycle comprises a vapor compressor and two independent ejectors operatively connected to high and low-pressure sides of the compressor, respectively. The two ejectors reduce the overall pressure ratio of the mechanical vapor compressor resulting in dramatically increased thermodynamic cycle efficiency. As one example of its potential applications for residential, commercial or industrial uses, a 150 ton capacity of a water-cooled chiller designed in accordance with the present invention is predicted to provide the power consumption as low as 0.47 kW/ton, when operated in accordance with the cooling methods of the present invention, which corresponds to 7.47 of Coefficient of Performance (COP).

Description

CROSS-REFERENCE TO RELATED APPLICATION

This application is a Continuation-in-Part of, and claims priority to, U.S. patent application Ser. No. 12/813,079 entitled “High Efficiency R744 Refrigeration System and Cycle,” filed on Jun. 10, 2010, which claims the benefit of U.S. Provisional Application No. 61/185,834, filed Jun. 10, 2009, entitled “High Efficiency R744 Refrigeration System and Cycle”, the aforementioned application being hereby incorporated by reference in its entirety.

TECHNICAL FIELD OF THE INVENTION

The present invention relates to refrigeration units and operation cycles, and more particularly to an R744 (carbon dioxide) vapor compression refrigeration system usable in numerous applications with improved cycle efficiency without any detrimental environmental effect.

BACKGROUND OF THE INVENTION

From residential or automobile air conditioning units to large packaged water chillers for air conditioning and refrigeration in commercial and industrial facilities, all mechanical vapor compression cycles now generally use environmentally harmful chemical refrigerants of hydrochlorofluorocarbon (HCFC) or hydrofluorocarbon (HFC) fluids. Emissions of halogenated synthetic refrigerants represent a major challenge for the environment due to their green house warming potential, as well as their contribution to ozone depletion. These harmful refrigerants are scheduled to be phased out, especially, HCFC-123, an ozone-depleting chemical, that has been designated by the Montreal Protocol and the U.S. Clean Air Act as a transitional refrigerant to be phased out soon. Similarly, the European Commission issued a directive in 2006 mandating the phase out of HFC-134a in mobile air conditioning units to help meet the European Union's Kyoto Protocol. Despite negligible ozone-depleting potential, HFC-134a has a very high Global Warming Potential (GWP) of 1,300 relative to 1 for natural refrigerants such as ammonia or carbon dioxide. Accordingly, the European Commission has instituted a requirement that mobile air conditioning units utilize refrigerants with a GWP no higher than 150.
A new synthetic refrigerant, HFO-1234yf, has been recently developed by DuPont and Honeywell for use in various air conditioning and refrigeration units. HFO-1234yf has a low GWP of about 4 to 6, and offers comparable performance to HFC-134a. However, there is still uncertainty about toxicity and flammability, which are critical to human life and safety. Suppliers say that HFO-1234yf has low acute and chronic toxicity, and manageable mild flammability; it still has a potential for A2 class of ISO 817 refrigerant classification, which means low toxicity and low flammability. However, in a press release issued on Oct. 31, 2008, Toyota Germany decided to move away from HFO-1234yf for car air conditioners after several independent tests raised issues about its flammability. The tests revealed that HFO-1234yf is flammable, releasing the decomposition product hydrogen fluoride that develops into the highly toxic gas hydrofluoric acid when in contact with water.
An alternative refrigerant option for air conditioning and refrigeration units has been to use R744 (carbon dioxide). R744 is a natural refrigerant that is free from ozone-depletion and has a negligible global warming potential. R744 has been used as a refrigerant for well over a century like other natural refrigerants, including R717 (ammonia), some of which have been excluded from use because of safety implications both in terms of toxicity and moderate flammability. R744 is neither flammable nor toxic, provided it is used in reasonable volumes, and thus provides desirable safety without having a detrimental environmental impact. However, in ordinary use, the low critical temperature of R744 (approximately 31° C.) forces the air conditioning system to work in the transcritical cycle at significantly higher pressures than desired. Accordingly, such prior art systems using R744 tend to utilize increased amounts of energy and thus, have not operated efficiently. For example, even though the increased performance density of R744 leads to smaller and lighter components, the basic transcritical cycle is potentially less efficient than a conventional vapor compression cycle because it suffers from larger thermodynamic losses at such a low critical temperature. Higher heat rejection temperatures with increased loss of energy result in greater throttling losses. As a result, the theoretical cycle work increases and refrigerant capacity is reduced. A direct application of R744 to the conventional air conditioning and refrigeration cycle therefore requires high power consumption, which is not desirable.
A simple prior art refrigeration cycle is generally illustrated in FIG. 1. As shown, the refrigeration cycle 10 includes a refrigerant loop through which a refrigerant is cycled, in order, through a compressor 12, a condenser 14, an expansion valve 16 and an evaporator 18. More particularly, a refrigerant enters the evaporator 18 as a mixture of liquid and vapor by being metered through the expansion valve 16, which lowers the pressure of the refrigerant, and therefore its temperature as well. Since the temperature of the refrigerant is colder than chilled water cycled through the evaporator 18, the refrigerant absorbs heat to boil into a saturated vapor. In order to dump out the absorbed heat, the refrigerant's temperature is raised by increasing its pressure using the compressor 12. In the condenser 14, the superheated vapor at the exit of the compressor 12 is condensed into liquid by losing heat into cooling water cycled through the condenser 14. The saturated liquid enters the expansion valve 16 to complete the cycle. The basic refrigeration cycle illustrated in FIG. 1 is currently utilized in most commercial air conditioning refrigeration units used today. Depending on the cooling loop fluid, the cycle may utilize either an air-cooled or a water-cooled condenser.
An alternate prior art refrigeration cycle is generally illustrated in FIG. 2. This refrigeration cycle has been recently developed for air conditioning units in passenger cars, and utilizes R744 refrigerant. As shown, the refrigeration cycle 110 includes an ejector 112, a separator 114, an internal heat exchanger 116, a compressor 118, a gas cooler 120, a metering valve 122, and an evaporator 124. In such a transcritical R744 cycle, a primary liquid flow with high pressure is supplied to the ejector 112 from the internal heat exchanger 116, and a secondary vapor flow with low pressure is supplied from the evaporator 124. The high-pressure and low-pressure flows are mixed in the ejector 112 and discharged at an intermediate pressure that is typically higher than the secondary vapor pressure. The mixed refrigerant is separated into a liquid flow and a vapor flow by the separator 114. The liquid flow separated by the separator 114 is throttled through the metering valve 122 and to the evaporator 124 to absorb heat. The vapor flow separated by the separator 114 is directed to the internal heat exchanger 116 to contribute to sub-cooling the ejector primary flow before entering the compressor 118. The gas cooler 120 is used to condense the compressed vapor flow into a liquid flow. A gas cooler is typically used in lieu of a condenser because of the use of supercritical R744 refrigerant. As noted above, the low critical temperature of R744 (approximately 31° C.) forces refrigeration cycle 110 to work in the transcritical cycle at significantly higher pressures than desired. Thus, the refrigeration cycle 110 requires high power consumption, which is not desirable or efficient.
In view of the foregoing, there is a need for a refrigeration unit and cycle that operates with high thermodynamic cycle efficiency with minimal harm to human safety and with minimal detrimental effect on the environment. Moreover, such a refrigeration unit and cycle should not be restricted to particular refrigerants in order to operate as desired. Indeed, a desirable system will operate efficiently regardless of the refrigerant used.
Additionally, there is a need for significant research efforts on the components and means used for refrigeration units and systems so as to improve cycle efficiency, including development and improvement of expanders (instead of expansion valves), ejectors and internal heat exchangers, so that losses incurred during a refrigeration cycle and operation of the system can be recovered.
Accordingly, it is a general object of the present invention to provide a refrigeration unit that overcomes the problems and drawbacks associated with existing refrigeration units and cycles and with use of various refrigerants.

SUMMARY OF THE INVENTION

The present invention is directed to an innovative air conditioning and refrigeration system. In a first aspect of the present invention, a refrigeration system using a refrigerant cycled therethrough, comprises a vapor compressor having a low-pressure side and a high-pressure side, a first ejector in operative communication with the high-pressure side of the vapor compressor, and a second ejector in operative communication with the low-pressure side of the vapor compressor. The first ejector boosts the pressure of vapor-phase refrigerant received from the vapor compressor using a pressurized sub-cooled liquid mixed with the vapor-phase refrigerant. The first ejector discharges a vapor stream having an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid. The second ejector boosts the pressure of vapor-phase refrigerant using sub-cooled liquid-phase refrigerant mixed with the vapor-phase refrigerant. Boosted vapor-phase refrigerant discharged from the second ejector is provided to the low-pressure side of the vapor compressor. The refrigeration system further comprises a condenser for converting the vapor stream discharged from the first ejector into a liquid-phase refrigerant, and a heat exchanger for converting the liquid-phase refrigerant into a sub-cooled liquid-phase refrigerant that is supplied to the first ejector, the second ejector, and preferably both ejectors.
The refrigeration system of the present invention may further comprise a centrifugal pump for pressurizing at least a portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger, wherein said pressurized sub-cooled liquid is provided to the first ejector for mixing with vapor-phase refrigerant therein.
In a second aspect of the present invention, a refrigeration system comprises a two-stage vapor compressor having a first compression stage and a second compression stage, the second compression stage operating at a higher pressure than the first compression stage. A first high-pressure ejector is in operative communication with the second compression stage of the vapor compressor, wherein the ejector boosts the pressure of vapor-phase refrigerant received from the vapor compressor using a pressurized sub-cooled liquid mixed with the vapor-phase refrigerant. The first ejector discharges a vapor stream having an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid. A condenser converts the vapor stream discharged from the first ejector into a liquid-phase refrigerant, and a heat exchanger converts the liquid-phase refrigerant into a sub-cooled liquid-phase refrigerant. A centrifugal pump pressurizes a first portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger, and the pressurized sub-cooled liquid is provided to the first ejector for mixing with vapor-phase refrigerant therein.
The refrigeration system also includes a second low-pressure ejector in operative communication with the first compression stage of the vapor compressor. The second ejector boosts the pressure of vapor-phase refrigerant using a second portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger that is mixed with the vapor-phase refrigerant in the second ejector.
In accordance with an advantage of the present invention, two independent ejectors are utilized to reduce total compression of the refrigeration cycle, and, as a result, reduce the electric power consumption of the vapor compressor. As a further result, the refrigeration system operates with higher cycle efficiency than for prior art systems.
In accordance with another advantage of the present invention, the system can be operated with negligible global warming potential and zero ozone depletion potential.
In accordance with yet another advantage of the present invention, the system is not limited to being used with any one refrigerant, which means that any refrigerants, including the preferred R744 refrigerant, may be used without departing from the spirit and principles of the invention.
In another aspect of the present invention a refrigeration cycling method comprises providing a mechanical vapor compressor having a low-pressure side and a high-pressure side, a first ejector in operative communication with the high-pressure side of the vapor compressor, and a second ejector in operative communication with the low-pressure side of the vapor compressor. Pursuant to the method, vapor-phase refrigerant discharged from the vapor compressor is mixed with a pressurized sub-cooled liquid in the first ejector so as to boost the pressure of the vapor-phase refrigerant to an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid. The vapor stream discharged from the first ejector is converted into a sub-cooled liquid-phase refrigerant. Within the second ejector, the vapor-phase refrigerant is mixed with the sub-cooled liquid-phase refrigerant so as to boost the pressure of the vapor-phase refrigerant to an elevated pressure. The mixed refrigerant discharged from the second ejector is then evaporated into a vapor-phase refrigerant that is provided through the heat exchanger to the mechanical vapor compressor. The vapor-phase refrigerant is also superheated and thereafter provided to the low-pressure side of the vapor compressor.
As one example of potential applications for the present invention, a 150-ton capacity, water-cooled chiller cycle is considered. The water-cooled chiller cycle is predicted through a preliminary cycle analysis for standard rating conditions controlled by ARI 550/590 to provide power consumption as low as 0.47 kW/ton, which corresponds to about 7.47 of Coefficient of Performance (COP).
The present invention cycle can be applied to air conditioning and refrigeration units and systems for various applications, including residential, automobile, industrial and commercial applications. The present invention is adaptable to various types of mechanical vapor compressors to accommodate various such applications. The liquid centrifugal pump of the disclosed embodiments can be either positive displacement machines or high-speed turbomachines. Heat exchanging methods used in the refrigeration cycle of the units or systems can be air-cooled, water-cooled or both.
These and other feature of the present invention are described with reference to the drawings of preferred embodiments of a refrigeration system. The illustrated embodiments of the system in accordance with the present invention are intended to illustrate, but not limit, the invention.

BRIEF DESCRIPTION DRAWINGS

FIG. 1 is a schematic diagram of a prior art refrigeration system and cycle.

FIG. 2 is a schematic diagram of another prior art refrigeration system and cycle.

FIG. 3 is a schematic diagram of a refrigeration system and cycle in accordance with the present invention.

FIG. 4 is an exemplary pressure-enthalpy thermodynamic diagram for a refrigeration cycle in accordance with the refrigeration system of FIG. 3 as used for air conditioning applications.

FIG. 5 is a schematic diagram of an alternative refrigeration system and cycle in accordance with the present invention.

FIG. 6 is an exemplary pressure-enthalpy thermodynamic diagram for a refrigeration cycle in accordance with the refrigeration system of FIG. 5 as used for air conditioning applications.

DETAILED DESCRIPTION OF THE INVENTION AND PREFERRED EMBODIMENTS THEREOF

FIGS. 3 and 4 represent an embodiment of a vapor compression refrigeration system and cycle in accordance with the present invention. More specifically, FIG. 3 illustrates the present invention in the form of a water-cooled chiller, generally designated by reference numeral 210, which is one of several applications for the innovative refrigeration cycle of the present invention. FIG. 4 provides an exemplary pressure-enthalpy thermodynamic diagram for the water-cooled chiller of FIG. 3. Both FIGS. 3 and 4 identify and correspond to various steps in the refrigeration cycle by reference to reference designators.
The illustrated water-cooled chiller generally has a 150-ton capacity and comprises a mechanical vapor compressor 212 driven by a high-speed electric motor 214. The vapor compressor 212 may be either a single stage or multi-staged compressor and of either a positive displacement-type or a centrifugal turbo-type. As illustrated in FIG. 3, the compressor 212 is a centrifugal, two-staged design, with a first compression stage 216 [Step B] and a second compression stage 218 [Step C]. In a design of the present invention utilizing a single-stage compressor, the compressor still comprises a high-pressure side and a low-pressure side. In preferred designs of the compression stages 216 and 218, each stage comprises an independent rotating impeller driven by the motor 214. In accordance with operation of the vapor compressor, as described below, the second compression stage 218 operates at a higher pressure than the first compression stage 216.
A high-pressure ejector 220 and a low-pressure ejector 230 are operatively connected to a high-pressure side and a low-pressure side of the vapor compressor 212, respectively. The high-pressure ejector 220 is preferably a two-phase jet device in which a sub-cooled liquid refrigerant, preferably R744 refrigerant, pressurized by a centrifugal pump 226 driven by a high-speed electric motor 228, is mixed with vapor refrigerant, compressed by the centrifugal vapor compressor 212. The centrifugal pump 226 can be either a positive displacement machine or a high-speed turbomachine.
As shown in FIG. 3, the compressed vapor refrigerant is discharged from the second compression stage 212 and delivered to a first input of the high-pressure ejector 220. The pressurized sub-cooled liquid refrigerant from the centrifugal pump 226 is delivered to a second input of the high-pressure ejector 220. The sub-cooled pressurized liquid refrigerant is used as a driving fluid to boost the vapor pressure of the vapor refrigerant received from the mechanical vapor compressor 212.
In operation, the high-pressure ejector 220 produces a final gas stream [Step F] with an elevated pressure that is higher than the pressure of either of the two inlet streams [Steps D and Q]. The high-pressure ejector 220 comprises a convergent section 220 a, a constant section 220 b and a diffuser section 220 c. The mixing of the pressurized sub-cooled liquid refrigerant with the vapor refrigerant takes place first in the convergent section 220 a of the ejector 220, and then in the constant section 220 b [Step R]. The two-phase mixed flow is converted into a gas stream across the R744 critical point in the diffuser section 220 c [Step E], as shown in FIG. 4. The high-pressure ejector 220 reduces total compression of the refrigeration cycle 210, and, as a result, the electric power consumption of the vapor compressor 212 is reduced. Such consumption levels are reduced, in part, by using the high-pressure ejector 220 in the manner described above to expand a portion of sub-cooled liquid refrigerant that is pressurized in advance by the centrifugal pump 226.
The gas stream ejected from an outlet of the high-pressure ejector 220 [Step F] is converted into a liquid refrigerant by a water-cooled condenser 222. As illustrated, the condenser 222 elevates the input water from about 85° F. to an output temperature of about 95° F. The liquid refrigerant is then fed to a secondary internal heat exchanger 224 [Step G], which converts the refrigerant into a sub-cooled liquid refrigerant. More particularly, the heat exchanger 224 allows the superheating process to take heat from the liquid refrigerant. Heat exchanging methods used in the refrigeration cycle of the present invention can be air-cooled, water-cooled or both. A portion of the sub-cooled liquid refrigerant, split at a flow branch 225 [Step H], is then depressurized and expanded through a nozzle arrangement at a first input of the low-pressure ejector 230. Another portion of the sub-cooled liquid from the secondary heat exchanger 224 is pressurized by the centrifugal pump 226. As noted above, the pressurized liquid from the pump 226 is utilized as a driving fluid in the high-pressure ejector 220 [Step Q] to boost vapor pressure from the mechanical vapor compressor 212.
The low-pressure ejector 230 is preferably a two-phase jet device comprising a convergent section 230 a, a constant section 230 b and a diffuser section 230 c. High-pressure refrigerant discharged from the low-pressure ejector 230 [Step K] exerts a drawing force to draw in a vapor-phase refrigerant [Step N], which is evaporated in an evaporator 236, through a second input of the low-pressure ejector 230. The evaporator 236 decreases the temperature of an input processing water from about 54° F. to an output temperature of about 44° F. The expansion energy of the refrigerant from the inlet of the low-pressure ejector 230 through the convergent section 230 a of the ejector (e.g., from Step H to Step I) is utilized to contribute to the increase of the intake pressure of the first compression stage 216 of the vapor compressor 212. The low-pressure ejector 230 reduces total compression of the refrigeration cycle 210 and therefore the electric power consumption of the vapor compressor 212 is reduced by recovering energy from the main expansion process for a portion of sub-cooled liquid refrigerant supplied to the ejector 230. Within the low-pressure ejector 230, the sub-cooled pressurized liquid refrigerant boosts the vapor pressure of the vapor-phase refrigerant received from the evaporator 236.
Liquid-phase refrigerant, which is separated by a liquid-vapor separator 232 downstream of the low-pressure ejector 230, is supplied to the evaporator 236 through an expansion valve 234 [Steps L and M]. The remaining vapor-phase refrigerant is supplied to the vapor compressor 212 through the secondary internal heat exchanger 224 to complete the vapor compression cycle [Steps P and A]. Specifically, the saturated vapor separated from the liquid-vapor separator 232 is superheated through the secondary internal heat exchanger 224 by the sub-cooling process downstream of the condenser 222, as discussed above [Step G].
In the refrigeration cycle 210 of the present invention, the centrifugal vapor compressor 212 has a two- stage compression 216 and 218, and intercooling is accomplished in the water-cooled condenser 222. The R744 refrigerant cycled through the system is utilized as a coolant for the electric motors 214 and 228. The loads of the high-pressure ejector 220 and the low-pressure ejector 230 may be cooled by air, water or both.
Thermodynamic processes of the refrigeration cycle for the water-cooled chiller 210 are illustrated in FIG. 4, where the designated points match with the cycle step designations in FIG. 3. For example, at Step A, vapor phase refrigerant is supplied to the low-pressure side 216 of the vapor compressor 212. The refrigerant is compressed through the first compression stage and passed through the water-cooled condenser at Step B. The refrigerant is then compressed through the second compression stage on the high-pressure side 218 of the vapor compressor at Step C. At Step D, a compressed vapor refrigerant is output from the second compression stage 218 and delivered to a first input of the high-pressure ejector 220. The high-pressure ejector 220 also receives a sub-cooled pressurized liquid from the centrifugal pump 226 at Step Q.
The high-pressure ejector 220 utilizes the pressurized liquid from the pump 226 as a driving fluid to boost the vapor pressure of the refrigerant received from the mechanical vapor compressor 212. Within the high-pressure ejector 220, at Step R, the pressurized sub-cooled liquid refrigerant is mixed with the vapor refrigerant, first in the convergent section 220 a of the ejector 220, and then in the constant section 220 b. Then, at Step E, the two-phase mixed flow is converted into a gas stream in the diffuser section 220 c of the ejector 220. At Step F, the high-pressure ejector 220 outputs a final gas stream with an elevated pressure that is directed to the condenser 222, which elevates the temperature of processing water while cooling the gas stream.
At Step G, the condenser sends the liquid refrigerant to the heat exchanger 224, which converts the refrigerant into a sub-cooled liquid refrigerant by allowing the superheating process to draw heat from the liquid refrigerant. At Step H, the sub-cooled refrigerant is split, with one part being directed to the centrifugal pump 226 for pressurization and delivery to the input of the high-pressure ejector 220, and the other part being directed to the low-pressure ejector 230.
The sub-cooled liquid refrigerant is mixed with vapor-phase refrigerant within the low-pressure ejector 230, more particularly in the convergent section 230 a and the constant section 220 b, at Step I. At Step J, the two-phase mixed flow is a high-pressure refrigerant in the divergent section 230 c of the ejector 230, which is then discharged from the ejector 230 and directed to the liquid-vapor separator 232 at Step K. The high-pressure refrigerant in the divergent section 230 c exerts a drawing force to draw in the vapor-phase refrigerant from the evaporator 236 at Step N for mixing within the low-pressure ejector 230.
At Steps L and P, liquid-vapor separator 232 separates the two-phase refrigerant flow into liquid-phase refrigerant and a saturated vapor-phase refrigerant. The liquid-phase refrigerant is directed through an expansion valve 234 and to the evaporator 236 at Step M, where the refrigerant is evaporated. The saturated vapor-phase refrigerant is superheated through the heat exchanger 224 at Step P, and then directed back to the vapor compressor at Step A.
FIGS. 5 and 6 represent another embodiment of a refrigeration system and cycle in accordance with the present invention. More specifically, FIG. 5 illustrates the present invention in the form of a water-cooled chiller, generally designated by reference numeral 310, which is one of several applications for the innovative refrigeration cycle of the present invention. Components of the chiller 310 are numbered similarly to like components of the chiller 210, discussed above and as shown in FIG. 3. FIG. 6 provides an exemplary pressure-enthalpy thermodynamic diagram for the water-cooled chiller of FIG. 5. Both FIGS. 5 and 6 identify and correspond to various steps in the refrigeration cycle by reference to reference designators.
Referring to FIG. 5, the water-cooled chiller 310 includes a similar set-up as for the water-cooled chiller 210. For example, the illustrated water-cooled chiller 310 comprises a mechanical vapor compressor 312 driven by a high-speed electric motor 314 and, as shown, a first compression stage 316 and a second compression stage 318. The compressor 312 can be of any variety, for example, a single stage or multi-staged compressor, and of either a positive displacement-type or a centrifugal turbo-type, and a low-speed or high-speed compressor.
In operation of a preferable vapor compressor 312 of the present invention, vapor enters the compressor 312 [Step A] and passes through one or two stages of compression before exiting the compressor 312 at a higher pressure than it entered. Thus, in accordance with intended operation of the present invention, as described below, the second compression stage 318 [Step C] operates at a higher pressure than the first compression stage 316 [Step B].
A high-pressure ejector 320 and a low-pressure ejector 330 are operatively connected to a high-pressure side and a low-pressure-side of the vapor compressor 312, respectively. The water-cooled chiller 310 generally includes the same high-pressure components as discussed above with respect to the water-cooled chiller 210 illustrated in FIG. 3, which are also arranged in a substantially similar manner. Unlike the water-cooled chiller 210, which includes a liquid-vapor separator 232 and a single evaporator 236, the water-cooled chiller 310 illustrated in FIG. 5 includes a primary evaporator 336 and a secondary evaporator 338 on the low-pressure side of the system, and utilizes no liquid-vapor separator.
As shown in FIG. 5, compressed vapor refrigerant is discharged from the second compression stage 318 and delivered to a first input—e.g., a suction port—of the high-pressure ejector 320 [Step D]. Pressurized sub-cooled liquid refrigerant from a centrifugal pump 326 is also delivered to the high-pressure ejector 320 via a second input [Step Q]. The high-pressure refrigerant vapor is mixed with the higher-pressure, sub-cooled liquid refrigerant in the high-pressure ejector 320, first in a convergent section 320 a and then in a constant section 320 b of the ejector 320 [Step R]. The sub-cooled pressurized liquid refrigerant is used as a driving fluid to boost the vapor pressure of the vapor refrigerant. In operation, the two-phase mixed flow is converted into a gas stream across the R744 critical point in a diffuser section 320 c of the ejector 320 [Step E], as shown in FIG. 6. Through the diffuser section 320 c of the ejector 320, the mixed flow recovers pressure energy so as to produce a final gas stream [Step F] with an elevated pressure that is higher than the pressure of either of the two inlet streams [Steps D and Q].
The high-pressure ejector 320 reduces total compression of the refrigeration cycle 310, and, as a result, the electric power consumption of the vapor compressor 312 is reduced. Such consumption levels are reduced, in part, by using the high-pressure ejector 320 in the manner described above to expand a portion of sub-cooled liquid refrigerant that is pressurized in advance by the centrifugal pump 326.
The gas stream ejected from an outlet of the high-pressure ejector 320 [Step F] is converted into a liquid refrigerant by a water-cooled condenser 322. As illustrated, the condenser 322 elevates the input water from about 85° F. to an output temperature of about 95° F. The liquid refrigerant is then fed to a secondary internal heat exchanger 324 [Step G], which converts the refrigerant into a sub-cooled liquid refrigerant. More particularly, the heat exchanger 324 allows the superheating process to take heat from the liquid refrigerant. As discussed above, heat exchanging methods used in the refrigeration cycle of the present invention can be air-cooled, water-cooled or both.
A portion of the sub-cooled liquid refrigerant, split at a flow branch 325 [Step H], is pressurized by the centrifugal pump 326. As noted above, the pressurized liquid from the pump 326 is utilized as a driving fluid in the high-pressure ejector 320 [Step Q] to boost vapor pressure from the mechanical vapor compressor 312. This refrigerant flow provides the suction energy necessary to entrain the vapor and boost its pressure upon exit from the high-pressure ejector 320.
Another portion of the sub-cooled liquid from the secondary heat exchanger 324, split at the flow branch 325 flows to another junction 340, where the flow is split again [Step T]. From flow junction 340, a portion of the sub-cooled liquid is then depressurized and expanded through a nozzle arrangement at a first inlet of the low-pressure ejector 230. The remaining portion of the sub-cooled liquid is directed through an expansion device 342 to undergo a pressure drop before entering the secondary evaporator 338 [Step M]. It is envisioned that the expansion device 342 can be a capillary tube, an expansion valve, or one of another variety of expansion device generally known in the art. In the secondary evaporator 338, the refrigerant absorbs heat from process water, evaporating at a constant temperature before exiting the secondary evaporator 338 as either a saturated or superheated vapor. From there, the refrigerant vapor is drawn through a second inlet—e.g., a suction port—of the low-pressure ejector 230 [Step O] and entrained by the motive flow—namely, the depressurized liquid drawn through the first inlet of the low-pressure ejector 330.
The low-pressure ejector 330 is preferably a two-phase jet device comprising a convergent section 330 a, a constant section 330 b and a diffuser section 330 c. High-pressure refrigerant discharged from the low-pressure ejector 330 [Step K] exerts a drawing force to draw in the vapor-phase refrigerant that is evaporated in the secondary evaporator 338 through the second inlet of the low-pressure ejector 330 [Step O]. The expansion energy of the refrigerant from the inlet of the low-pressure ejector 330 through the convergent section 330 a of the ejector (e.g., from Step H to Step I) is utilized to contribute to the increase of the intake pressure of the first compression stage 316 of the vapor compressor 312. The low-pressure ejector 330 reduces total compression of the refrigeration cycle 310 and therefore the electric power consumption of the vapor compressor 312 is reduced by recovering energy from the main expansion process for a portion of sub-cooled liquid refrigerant supplied to the ejector 330. Within the low-pressure ejector 330, the sub-cooled pressurized liquid refrigerant boosts the vapor pressure of the vapor-phase refrigerant received from the secondary evaporator 338. Thus, just as in the high-pressure ejector 320, the refrigerant recovers pressure through the diffuser section 330 c of the low-pressure ejector 330 and exits at a higher pressure than that of the inlet vapor.
The mixed-phase refrigerant flows out of the low-pressure ejector 330 and enters the primary evaporator 336 [Step K]. In the primary evaporator 336, the refrigerant absorbs heat from the process water and evaporates. Upon exiting from the primary evaporator 336 [Step P], the refrigerant is either a saturated or superheated vapor. The vapor-phase refrigerant then passes through the secondary internal heat exchanger 324, subcooling the liquid from the water-cooled condenser 322 and superheating the vapor-phrase refrigerant prior to entering the first compression stage 316 [Step A].
The process water used for the evaporators 336 and 338 first enters the primary evaporator 336 at a temperature of about 54° F. In the primary evaporator 336, the refrigerant preferably evaporates at a higher temperature and pressure than normal due to compression through the low-pressure ejector 330. The process water rejects heat to the evaporating refrigerant before exiting the primary evaporator 336 and entering the secondary evaporator 338. In the secondary evaporator 338, the refrigerant preferably evaporates at a more standard temperature that is lower than the temperature in the primary evaporator 336. In the secondary evaporator 338, the process water rejects additional heat to the evaporating refrigerant before exiting the secondary evaporator 338, and is chilled to the desired output temperature of about 44° F. In accordance with an aspect of the present invention, the secondary evaporator 338 is smaller than the primary evaporator 336.
In operation, the mixed-phase refrigerant that passes through the condenser 322 rejects heat that was gained on the low-pressure side of the system by the primary and secondary evaporators 336 and 338.
Thermodynamic processes of the refrigeration cycle for the water-cooled chiller 310 are illustrated in FIG. 6, where the designated points match with the cycle step designations noted in FIG. 5. For example, at Step A, vapor phase refrigerant is supplied to the low-pressure side 316 of the vapor compressor 312. The refrigerant is compressed through the first compression stage and passed through the water-cooled condenser at Step B. The refrigerant is then compressed through the second compression stage on the high-pressure side 318 of the vapor compressor at Step C. At Step D, a compressed vapor refrigerant is output from the second compression stage 318 and delivered to a first input of the high-pressure ejector 320. The high-pressure ejector 320 also receives a sub-cooled pressurized liquid from the centrifugal pump 326 at Step Q.
The high-pressure ejector 320 utilizes the pressurized liquid from the pump 326 as a driving fluid to boost the vapor pressure of the refrigerant received from the mechanical vapor compressor 312. Within the high-pressure ejector 320, at Step R, the pressurized sub-cooled liquid refrigerant is mixed with the vapor refrigerant, first in the convergent section 320 a of the ejector 320, and then in the constant section 320 b. Then, at Step E, the two-phase mixed flow is converted into a gas stream in the diffuser section 320 c of the ejector 320. At Step F, the high-pressure ejector 320 outputs a final gas stream with an elevated pressure that is directed to the condenser 322, which elevates the temperature of processing water while cooling the gas stream.
At Step G, the condenser 322 sends the liquid refrigerant to the heat exchanger 324, which converts the refrigerant into a sub-cooled liquid refrigerant by allowing the superheating process to draw heat from the liquid refrigerant. At Step H, the sub-cooled refrigerant is split, with one part being directed to the centrifugal pump 326 for pressurization and delivery to the input of the high-pressure ejector 320, and the other part being directed to the flow junction 340. At Step T, the refrigerant is again split, with one part being directed to the low-pressure ejector 330, and the other part being directed to the expansion device 342.
The sub-cooled liquid undergoes a pressure drop in the expansion device 342 and is directed into the secondary evaporator 338 at Step M. In the secondary evaporator 338, the refrigerant is evaporated by process water and exits at Step N as either a saturated or superheated vapor. From there, the refrigerant vapor enters the suction port of the low-pressure ejector 330 at Step O. The sub-cooled liquid refrigerant is mixed with vapor-phase refrigerant within the low-pressure ejector 330, more particularly in the convergent section 330 a and the constant section 320 b, at Step I. At Step J, the two-phase mixed flow is a high-pressure refrigerant in the divergent section 330 c of the ejector 330, which is then discharged from the ejector 330 and directed to the primary evaporator 336 at Step K. The high-pressure refrigerant in the divergent section 330 c exerts a drawing force to draw in the vapor-phase refrigerant from the secondary evaporator 338 at Step N for mixing within the low-pressure ejector 330.
In the primary evaporator 336, the high-pressure refrigerant absorbs heat from process water and evaporates. The flow exiting the primary evaporator 336 at Step P is either a saturated or superheated vapor. The vapor-phase refrigerant is superheated through the heat exchanger 324, and then directed back to the vapor compressor at Step A.
Both the high-pressure ejector and the low-pressure ejector used in various embodiments of the present invention reduce total compression of the refrigeration cycle and the power consumption of the vapor compressor. With regard to the high-pressure ejector, consumption levels are reduced, in part, by using the high-pressure ejector to expand a portion of sub-cooled liquid refrigerant that is pressurized by the centrifugal pump in advance of feeding the liquid refrigerant to the ejector. With regard to the low-pressure ejector, consumption levels are reduced by recovering energy from the main expansion process for a portion of sub-cooled liquid refrigerant.
Refrigeration systems and cycles in accordance with the present invention provides higher cycle efficiency without increases electric power consumption to undesirable levels. In particular, the design of the system reduces the overall pressure ratio of the vapor compressor, resulting in dramatically increased thermodynamic cycle efficiency. Additionally, the present invention permits a refrigeration system and cycle to operate with negligible global warming potential and zero ozone depletion potential. Moreover, the present invention is not limited to use with particular refrigerant, which means that any refrigerant may be utilized without compromising the efficiency of the system. R744 (carbon dioxide) is discussed as a preferred refrigerant because of its known negligible global warming potential and zero ozone depletion potential. However, other refrigerants may be utilized in the refrigeration system and cycle described herein without departing from the spirit and principles of the invention.
Based on standard rating conditions of packaged water chillers controlled by ARI 550/590, a preliminary evaluation of the innovative refrigeration cycle of the present invention for 150-ton capacity shows 0.47 kW/ton of power consumption and 7.47 of COP when both ejectors have a pressure ratio of about 1.2, and 81% and 80% of isentropic efficiencies are assumed for the first and second stages of a two-stage centrifugal vapor compressor. A 50% split of mass flow rate is also assumed at the flow branch 225 or 325 illustrated in the embodiments. In the case of 150 tons of cooling capacity, the centrifugal vapor compressor consumes about 68 kW of input power, and the centrifugal pump requires only about 2.5 kW. Considering the averaged COP level of current state-of-the-art water chillers using HFC-134a is around 5.5, the present innovative refrigeration cycle provides a great improvement in energy savings.
The foregoing description of the present invention has been presented for the purpose of illustration and description. It is not intended to be exhaustive as to limit the invention to the form disclosed. Obvious modifications and variations are possible in light of the above disclosure. The embodiments described were chosen to best illustrate the principles of the invention and practical applications thereof to enable one of ordinary skill in the art to utilize the invention in various embodiments and with various modifications as suited to the particular uses contemplated.
The present invention cycle can be applied to air conditioning and refrigeration units and systems for a variety of applications, including residential, automobile, industrial and commercial applications. The present invention is adaptable to various types of mechanical vapor compressors to accommodate various such applications. The liquid pump of the disclosed embodiments can be either positive displacement machines or high-speed turbomachines. Heat exchanging methods used in the refrigeration cycle of the units or systems can be air-cooled, water-cooled or both. FIG. 1 shows one of potential applications.
It is intended that the scope of the present invention be defined by the claims appended hereto.

Claims

1. A refrigeration system using a refrigerant cycled therethrough, said refrigeration system comprising:

a vapor compressor comprising a low-pressure side and a high-pressure side;

a first ejector in operative communication with the high-pressure side of the vapor compressor, wherein the first ejector boosts the pressure of vapor-phase refrigerant received from the vapor compressor using a pressurized sub-cooled liquid mixed with said vapor-phase refrigerant, and further wherein said first ejector discharges a vapor stream having an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid;

a condenser for converting the vapor stream discharged from the first ejector into a liquid-phase refrigerant;

a heat exchanger for converting the liquid-phase refrigerant into a sub-cooled liquid-phase refrigerant;

a second ejector in operative communication with the low-pressure side of the vapor compressor, the second ejector boosts the pressure of vapor-phase refrigerant using at least a portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger mixed with said vapor-phase refrigerant, and the refrigerant from the second ejector is provided through the heat exchanger for subcooling the liquid-phase refrigerant to the low-pressure side of the vapor compressor.

2. The refrigeration system as claimed in claim 1, further comprising:

a primary evaporator in operative communication with a discharge outlet of the second ejector for evaporating the mixed-phase refrigerant into a vapor-phase refrigerant; and

a secondary evaporator in operative communication with the heat exchanger for evaporating the sub-cooled liquid-phase refrigerant into a vapor-phase refrigerant that is provided to an input of the second ejector for mixing with said at least a portion of the sub-cooled liquid phase refrigerant therein.

3. The refrigeration system as claimed in claim 2, further comprising an expansion valve operatively positioned between the heat exchanger and the secondary evaporator for expanding the liquid-phase refrigerant provided to the secondary evaporator.

4. The refrigeration system as claimed in claim 2, wherein the vapor-phase refrigerant evaporated by the primary evaporator is superheated through the heat exchanger before being provided to the low-pressure side of the vapor compressor.

5. The refrigeration system as claimed in claim 1, further comprising a centrifugal pump for pressurizing at least a portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger, wherein said pressurized sub-cooled liquid is provided to the first ejector for mixing with vapor-phase refrigerant therein.

6. The refrigeration system as claimed in claim 1, wherein the vapor compressor is a single-stage compressor.

7. The refrigeration system as claimed in claim 1, wherein the vapor compressor is a two-stage compressor with a first compression stage and a second compression stage, said second compression stage operating at a higher pressure than the first compression stage.

8. The refrigeration system as claimed in claim 7, wherein each stage of the vapor compressor is driven by a motor.

9. The refrigeration system as claimed in claim 8, wherein each stage of the vapor compressor includes an impeller operatively connected to the motor for rotation.

10. The refrigeration system as claimed in claim 1, wherein the loads of the first and second ejectors are cooled by at least one of air or water.

11. The refrigeration system as claimed in claim 1, wherein the vapor compressor is a positive displacement compressor.

12. The refrigeration system as claimed in claim 1, wherein the vapor compressor is a centrifugal turbocompressor.

13. A refrigeration system using a refrigerant cycled therethrough, said refrigeration system comprising:

a two-stage vapor compressor comprising a first compression stage and a second compression stage, said second compression stage operating at a higher pressure than the first compression stage;

a first high-pressure ejector in operative communication with the second compression stage of the vapor compressor, wherein the first ejector boosts the pressure of vapor-phase refrigerant received from the vapor compressor using a pressurized sub-cooled liquid mixed with said vapor-phase refrigerant, and further wherein said first ejector discharges a vapor stream having an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid;

a centrifugal pump for pressurizing a first portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger, wherein said pressurized sub-cooled liquid is provided to the first ejector for mixing with vapor-phase refrigerant therein;

a second low-pressure ejector in operative communication with the first compression stage of the vapor compressor, wherein the second ejector boosts the pressure of vapor-phase refrigerant using a second portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger mixed with said vapor-phase refrigerant, and further wherein the second ejector discharges a mixed refrigerant;

a primary evaporator in operative communication with a discharge outlet of the second ejector for evaporating the mixed refrigerant discharged therefrom into a vapor-phase refrigerant; and

a secondary evaporator that evaporates a portion of the second portion of the sub-cooled liquid-phase refrigerant discharged from the heat exchanger into a vapor-phase refrigerant that is provided to an input of the second ejector for mixing with the remainder of said second portion of the sub-cooled liquid-phase refrigerant therein;

wherein the vapor-phase refrigerant discharged from the primary evaporator is provided to the first compression stage of the vapor compressor.

14. The refrigeration system as claimed in claim 13, further comprising an expansion valve operatively positioned between the heat exchanger and the secondary evaporator for expanding the liquid-phase refrigerant provided to the evaporator.

15. The refrigeration system as claimed in claim 14, wherein the vapor-phase refrigerant from the primary evaporator is superheated through the heat exchanger before being provided to the low-pressure side of the vapor compressor.

16. The refrigeration system as claimed in claim 14, wherein the loads of the first and second ejectors are cooled by at least one of air or water.

17. A refrigeration cycling method comprising:

providing a mechanical vapor compressor having a low-pressure side and a high-pressure side;

providing a first ejector in operative communication with the high-pressure side of the vapor compressor;

mixing vapor-phase refrigerant discharged from the vapor compressor with a pressurized sub-cooled liquid in the first ejector so as to boost the pressure of the vapor-phase refrigerant to an elevated pressure that is greater than the pressure of either the input vapor-phase refrigerant received from the vapor compressor or the pressurized sub-cooled liquid;

converting the vapor stream discharged from the first ejector into a sub-cooled liquid-phase refrigerant;

providing a second ejector in operative communication with the low-pressure side of the vapor compressor;

mixing vapor-phase refrigerant with the sub-cooled liquid-phase refrigerant in the second ejector to provide a mixed vapor-phase refrigerant having an elevated pressure;

evaporating the mixed-phase refrigerant into a vapor-phase refrigerant that is provided through the heat exchanger to the mechanical vapor compressor;

superheating the vapor-phase refrigerant; and

providing the superheated vapor-phase refrigerant to the low-pressure side of the vapor compressor.

18. The refrigeration cycling method as claimed in claim 17, wherein the vapor compressor is a two-stage compressor with a first compression stage and a second compression stage, said second compression stage operating at a higher pressure than the first compression stage.

19. The refrigeration cycling method as claimed in claim 17, wherein each of the first and second ejectors comprises a two-phase ejector.

20. The refrigeration cycling method as claimed in claim 17, further comprising the step of cooling the loads of the first and second ejectors by at least one of air or water.