EP0424474A1 - Method of operating a vapour compression cycle under trans- or supercritical conditions. - Google Patents

Method of operating a vapour compression cycle under trans- or supercritical conditions.

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Publication number
EP0424474A1
EP0424474A1 EP89910211A EP89910211A EP0424474A1 EP 0424474 A1 EP0424474 A1 EP 0424474A1 EP 89910211 A EP89910211 A EP 89910211A EP 89910211 A EP89910211 A EP 89910211A EP 0424474 A1 EP0424474 A1 EP 0424474A1
Authority
EP
European Patent Office
Prior art keywords
refrigerant
pressure
capacity
heat exchanger
temperature
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Granted
Application number
EP89910211A
Other languages
German (de)
French (fr)
Other versions
EP0424474B1 (en
EP0424474B2 (en
Inventor
Gustav Lorentzen
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
SINVENT A/S TE TRONDHEIM, NOORWEGEN.
Original Assignee
Elcraft AS
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Filing date
Publication date
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B45/00—Arrangements for charging or discharging refrigerant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B40/00—Subcoolers, desuperheaters or superheaters
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F25B41/20—Disposition of valves, e.g. of on-off valves or flow control valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B9/00—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point
    • F25B9/002—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant
    • F25B9/008—Compression machines, plants or systems, in which the refrigerant is air or other gas of low boiling point characterised by the refrigerant the refrigerant being carbon dioxide
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2309/00—Gas cycle refrigeration machines
    • F25B2309/06—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide
    • F25B2309/061—Compression machines, plants or systems characterised by the refrigerant being carbon dioxide with cycle highest pressure above the supercritical pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/04—Refrigeration circuit bypassing means
    • F25B2400/0411—Refrigeration circuit bypassing means for the expansion valve or capillary tube
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/04—Refrigeration circuit bypassing means
    • F25B2400/0415—Refrigeration circuit bypassing means for the receiver
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/16—Receivers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/17—Control issues by controlling the pressure of the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2600/00—Control issues
    • F25B2600/25—Control of valves
    • F25B2600/2501—Bypass valves

Definitions

  • This invention relates to vapour compression cycle devices such as refrigerators, air-conditioning units and heat pumps, using a refrigerant operating in a closed circuit under trans-critical conditions, and more particularly, to methods for modulating and controlling the capacity of such devices.
  • a conventional vapour compression cycle device for refri ⁇ geration, air-conditioning or heat pump purposes is shown in principle in Fig. 1.
  • the device consists of a compressor (1), a condensing heat exchanger (2), a throttling valve (3) and a evaporating heat exchanger (4). These components are connected in a closed flow circuit, in which a refrigerant is circulated.
  • the operating principle of a vapour compres ⁇ sion cycle device is as follows: The pressure and tempera ⁇ ture of the refrigerant vapour is increased by the compres ⁇ sor (1), before it enters the condenser (2) where it is cooled and condensed, giving off heat to a secondary cool ⁇ ant.
  • the high-pressure liquid is then throttled to the eva ⁇ porator pressure and temperature by means of the expansion valve (3).
  • the evaporator (4) the refrigerant boils and absorbs heat from its surroundings.
  • the vapour at the eva ⁇ porator outlet is drawn into the compressor, completing the cycle.
  • refrige ⁇ rants as for instance R-12, CF 2 C1 2
  • refrigerant A number of different substances or mixtures of substances may be used as a refrigerant.
  • the choice of refrigerant is among others influenced by the condensation temperature, as the critical temperature of the fluid sets the upper limit for the condensation to occur. In order to maintain a reasonable efficiency, it is normally desirable to use a refrigerant with critical temperature at least 20-30K above the condensation temperature. Near- critical temperatures are normally avoided in design and operation of conventional systems.
  • Capacity control of the conventional vapour compression cycle device is achieved mainly by regulating the mass flow of refrigerant passing through the evaporator. This is done e.g. by controlling the compressor capacity, throttling or bypassing. These methods involve more complicated flow cir ⁇ cuit and components, need for additional equipment and ac- cesories, reduced part-load efficiency and other compli ⁇ cations.
  • a common type of liquid regulation device is a thermostatic expansion valve, which is controlled by the superheat at the evaporator outlet. Proper valve operation under varying operating conditions is achieved by using a considerable part of the evaporator to superheat the refrigerant, resul ⁇ ting in a low heat transfer coefficient.
  • thermodynamic losses occur due to large temperature differences when giving off heat to a secondary coolant with large temperature increase, as in heat pump applications or when the available secondary cool ⁇ ant flow is small.
  • Another object of the present invention is to provide a vapour compression cycle avoiding use of CFC refrigerants, and at the same time offering possibility to apply several attractive refrigerants with respect to safety, environ ⁇ mental hazards and price.
  • Further object of the present invention is to provide a new method of capacity control, which involves operation at mainy constant refrigerant mass flow rate and simple capa ⁇ city modulation by valve operation.
  • Still another object of the present invention is to provide a cycle rejecting heat at gliding temperature, reducing heat-exchange losses in applications where secondary coolant flow is small or when the secondary coolant is to be heated to a relatively high temperature.
  • thermodynamic properties in the super-critical state are utilized to con ⁇ trol the refrigerating and heating capacity of the device.
  • the present invention involves the regulation of specific enthalpy at evaporator inlet by deliberate use of the pres ⁇ sure and/or temperature before throttling for capacity con ⁇ trol. Capacity is controlled by varying the refrigerant enthalpy difference in the evaporator, by changing the spe ⁇ cific enthalpy of the refrigerant before throttling. In the super-critical state this can be done by varying the pres ⁇ sure and temperature independently. In a preferred embodi ⁇ ment this modulation of specific enthalpy is done by varying the pressure before throttling.
  • the refrigerant is cooled down as far as it is feasible by means of the available cooling medium, and the pressure regulated to give the re ⁇ quired enthalpy.
  • Another embodiment involves modulation of enthalpy by variation of the refrigerant temperature before throttling. This is done by controlling the heat rejection from the device.
  • Fig. 1 is a schematic representation of a conventional (sub- critical) vapour compression cycle device.
  • Fig. 2 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a preferred embodiment of the invention.
  • This embodi ⁇ ment includes a volume as an integral part of the evaporator system, holding refrigerant in the liquid state.
  • Fig. 3 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a second embodiment of the invention.
  • This embodiment includes an intermediate pressure receiver connected direct ⁇ ly into the flow circuit between two valves.
  • Fig. 4 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a third embodiment of the invention.
  • This embodiment includes a special receiver to hold refrigerant as liquid or in the super-critical state.
  • Fig. 5 is a graph illustrating the relationship of pressure versus enthalpy of the trans-critical vapour compression cycle device of Fig. 2, 3 or 4, at different operating con ⁇ ditions.
  • Fig. 6 is a collection of graphs illustrating the control of refrigerating capacity by the method of pressure control in accordance with the present invention. The results shown are measured in a laboratory demonstration system built accord ⁇ ing to a preferred embodiment of the invention.
  • Fig. 7 is a collection of graphs illustrating the control of refrigerating capacity by control of the heat rejection, in accordance with the present invention. The results shown are measured in a laboratory demonstration system build accord ⁇ ing to a preferred embodiment of the present invention.
  • Fig. 8 is test results showing the relationship of tempera ⁇ ture versus entropy of the trans-critical vapour compression cycle device of Fig. 2, operating at different high-side pressures, employing carbon dioxide as refrigerant Detailed description of the invention
  • a trans-critical vapour compression cycle device includes a refrigerant, of which critical temperature is between the temperature of the heat inlet and the mean temperature of heat submittal, and a closed working fluid circuit where the refrigerant is circu ⁇ lated.
  • Suitable working fluids may be by the way of examples: ethy- len (C 2 H 4 ), diborane (B 2 H 6 ), carbon dioxide (C0 2 ), ethane (C 2 H 6 ) and nitrogen oxide (N 2 0).
  • the closed working fluid circuit consists of a refrigerant flow loop with an integrated storage segment.
  • Fig. 2 shows a preferred embodiment of the invention where the storage segment is an integral part of the evaporator system.
  • the flow circuit includes a compressor 10 connected in series to a heat exchanger 11, a counterflow heat exchanger 12 and a throttling valve 13.
  • the throttling valve can be replaced by an optional expansion device.
  • An evaporating heat exchanger 14, a liquid separator/receiver 16 and the low-pressure side of the counterflow heat exchanger 12 are connected in flow communication intermediate the throttling valve 13 and the inlet 19 of the compressor 10.
  • the liquid receiver 16 is connected to the evaporator outlet 15, and the gas phase outlet of the receiver 16 is connected to the counterflow heat exchanger 12.
  • the counterflow heat exchanger 12 is not absolutely neces ⁇ sary for the functioning of the device but improves its efficiency, in particular its rate of response to a capacity increase requirement. It also serves to return oil to the compressor.
  • a liquid phase line from the receiver (16) (shown with broken line in Fig. 2) is connect ⁇ ed to the suction line either before the counterflow heat exchanger (12) at 17 or after it at 18, or anywhere between these points.
  • the liquid flow i.e. refrigerant and oil, is controlled by a suitable conventional liquid flow restrict ⁇ ing device (not shown in the figure). By allowing some ex- cess liquid refrigerant to enter the vapour line, a liquid surplus at the evaporator outlet is obtained.
  • the storage segment of the working fluid circuit includes a receiver 22 integrated in the flow circuit between a valve 21 and the throttling valve 13.
  • the other components 10-14 of the flow circuit is identical to the components of the previous embodiment, although the heat exchanger 12 can be omitted without any great consequence.
  • the pressure in the receiver 22 is kept intermediate the high-side and low-side pressures of the flow circuit.
  • the storage segment of the working fluid circuit includes a special receiver 25, where the pressure is kept between the high-side pressure and the low-side pressure of the flow circuit.
  • the storage segment further consists of the valves 23 and 24 which are connected to the high pressure and low pressure part of the flow circuit respectively.
  • the refrigerant is compressed to a suitable supercritical pressure in the compressor 10, the compressor outlet 20 is shown as state “a” in Fig. 5.
  • the refrigerant is circulated through the heat exchanger 11 where it is cooled to state "b", giving off heat to a suitable cooling agent, e.g. cooling air or water.
  • a suitable cooling agent e.g. cooling air or water.
  • the refrig ⁇ erant can be further cooled to state "c" in the counterflow heat exchanger 12, before throttling to state "d".
  • a two-phase gas/liquid mixture is formed, shown as state “d” in Fig. 3.
  • the refrigerant absorbs heat in the evaporator 14 by eva ⁇ poration of the liquid phase.
  • the refrigerant vapour can be superheated in the counterflow heat exchanger 12 to state “f" before it enters the compressor inlet 19, making the cycle complete.
  • the evaporator outlet condition "e” will be in the two-phase region due to the liquid surplus at the evaporator outlet. Modulation of the trans-critical cycle device capacity is accomplished by varying the refrigerant state at the eva ⁇ porator inlet, i.e. point "d" in Fig. 5.
  • the refrigerating capacity per unit of refrigerant mass flow corresponds to the enthalpy difference between state “d” and state “e”; This enthalpy difference is found as a horizontal distance in the enthalpy-pressure diagram. Fig. 5.
  • Throttling is a constant enthalpy process, thus the enthalpy in point “d” is equal to the enthalpy in point "c".
  • the ref igerating capacity (in k ) at constant refrigerant mass flow can be controlled by varying the en ⁇ thalpy at point "c".
  • the high-pressure single-phase refrigerant vapour is not conden ⁇ sed but reduced in temperature in the heat exchanger 11.
  • the terminal temperature of the refrigerant in the heat ex ⁇ changer (point “b” ) will be some degrees above the entering cooling air or water temperature, if counterflow is used.
  • the high-pressure vapour can then be cooled a few degrees lower, to point "c", in the counterflow heat exchanger 12. The result is, however, that at constant cooling air or water inlet temperature, the temperature at point "c" will be mainly constant, independent of the pressure level in the high side.
  • modulation of device capacity is accomplished by varying the pressure in the highside, while the temperature in point "c" is mainly constant.
  • the curvature of the iso ⁇ therms near the critical point result in a variation of enthalpy with pressure, as shown in Fig. 5.
  • the figure shows a reference cycle (a-b-c-d-e-f) , a cycle with reduced capa ⁇ city due to reduced high side pressure (a'-b'-c'-d'-e-f) and a cycle with increased capacity due to higher pressure in the high side (a"-b"-c"-d"-e-f) .
  • the evaporator pressure is assumed to be constant.
  • the pressure in the high-pressure side is independent of temperature, because it is filled with a single phase fluid. To vary the pressure it is necessary to vary the mass of refrigerant in the high side, i.e. to add or remove some of the instant refrigerant charge in the high side. These vari ⁇ ations must be taken up by a buffer, to avoid liquid over ⁇ flow or drying up of the evaporator.
  • the refrigerant mass in the high side can be in ⁇ creased by temporarily reducing the opening of the throt ⁇ tling valve 13. Due to the incidentally reduced refrigerant flow to the evaporator, the excess liquid fraction at the evaporator outlet (15) will be reduced. The liquid refri ⁇ gerant flow from the receiver 16 into the suction line is however constant. Consequently, the balance between the liquid flow entering and leaving the receiver 16 is shifted, resulting in a net reduction in receiver liquid content and a corresponding accumulation of refrigerant in the high pressure side of the flow circuit.
  • the increase in high side charge involves increasing pres ⁇ sure and thereby higher refrigerating capacity. This mass transfer from the low-pressure to the high-pressure side of the circuit will continue until balance between refrigerat ⁇ ing capacity and load is found.
  • Opening of the throttling valve 13 will increase the excess liquid fraction at the evaporator outlet 15, because the evaporated amount of refrigerant is mainly constant. The difference between this liquid flow entering the receiver and the liquid flow from the receiver into the suction line, will accumulate. The result is a net transport of refri ⁇ gerant charge from the high side to the low side of the flow circuit, with the reduction in the high side charge stored in liquid state in the receiver. By reducing the high-side charge and thereby pressure, the capacity of the device is reduced, until balance is found.
  • the refrigerant mass in the high side can be increased by simultaneously shutting the valve 21 and modulating the throttling valve 13 to provide the evaporator with suf ⁇ ficient liquid flow. This will reduce the refrigerant flow from the high side into the receiver through valve 21, while refrigerant mass is transferred from the low side to the high side by the compressor.
  • Reduction of high-side charge is obtained by opening the valve 21 while keeping the flow through the throttling valve 13 mainly constant. This will transfer mass from the high- side of the flow circuit to the receiver 22.
  • the refrigerant mass in the high side can be increased by opening the valve 24 and simultaneously reducing the flow through the throttling valve 13.
  • refrigerant charge is accumulated in the high-pressure side due to reduced flow through the throttling valve 13.
  • Sufficient liquid flow to the evaporator is obtained by opening the valve 24.
  • a reduction in the high side charge can be accomplished by opening the valve 23 to transfer some refrigerant charge from the high side to the receiver. Capacity control of the device is thus accomplished by modulation of the valves 23 and 24, and simultaneously operating the throttling valve 13.
  • the preferred embodiment of the invention has the advantage of simplicity, with capacity con ⁇ trol by operation of one valve only. Furthermore, the trans- critical vapour compression cycle device built according to this embodiment has a certain self-regulating capability by adapting to changes in cooling load through changes in liquid content in the receiver 16, involving changes in highside charge and thus cooling capacity. In addition, the operation with liquid surplus at evaporator outlet gives favourable heat transfer characteristics.
  • the second embodiment as indicated in Fig. 3, has the ad ⁇ vantage of simplified valve operation. Valve 21 only regu ⁇ lates the pressure in the high side of the device, and the throttling valve 13 only assures that the evaporator is fed sufficiently. A conventional thermostatic valve can thus be used for throttling.
  • Oil return to the compressor is easily achieved by allowing the refrigerant to flow through the receiver.
  • This embodiment does not offer the capa ⁇ city control function at high-side pressures below the cri ⁇ tical pressure.
  • the volume of the receiver 22 must be rela ⁇ tively large since it is only operating between the dis ⁇ charge pressure and the liquid line pressure.
  • Still another embodiment as indicated in Fig. 4 has the advantage of operating as a conventional vapour compression cycle device, when it is running at stable conditions.
  • the valves 23 and 24, connecting the receiver 25 to the flow circuit, are activated only during capacity control.
  • This embodiment requires use of three different valves during periods of capacity change.
  • Trans-critical vapour compression cycle devices built ac ⁇ cording to the described embodiments can be applied in seve ⁇ ral areas.
  • the technology is well suitable in small and medium-sized stationary and mobile air-conditioning units, small and medium-sized refrigerators/freezers and in smaller heat pump units.
  • One of the most promising applications is in automotive air-conditioning, where the present need for a new, non-CFC, lightweight and efficient alternative to R12- systems is urgent.
  • the laboratory test device uses water as heat source, i.e. the water is refrigerated by heat exchange with boiling CO 2 in the evaporator 14. Water is also used as cooling agent, being heated by C0 2 in the heat exchanger 11.
  • the test de ⁇ vice includes a 61 ccm reciprocating compressor (10) and a receiver (16) with total volume of 4 liters.
  • the system also includes a counterflow heat exchanger (12) and liquid line connection from the receiver to point 17, as indicated in Fig. 2.
  • the throttling valve 13 is operated manually.
  • This example shows how control of refrigerating capacity is obtained by varying the position of the throttling valve 13, thereby varying the pressure in the high-side of the flow circuit.
  • the specific refrigerant enthalpy at the evaporator inlet is controlled, resulting in modulation of refrigerating capacity at constant mass flow.
  • the water inlet temperature to the evaporator 14 is kept constant at 20°C, and the water inlet temperature to the heat exchanger 11 is kept constant at 35°C.
  • Water circu ⁇ lation is constant both in the evaporator 14 and the heat exchanger 11.
  • the compressor is running at constant speed.
  • Fig. 6 shows the variation of refrigerating capacity (Q), compressor shaft work (W), high ⁇ ide pressure (p H ), C0 2 mass flow (m), C0 2 temperature at evaporator outlet (t # ), C0 2 temperature at the outlet of heat exchanger 11 (t b ) and liquid level in the receiver (h) when the throttling valve 13 is operated as indicated at the top of the figure.
  • the adjustment of throttling valve position is the only mani ⁇ pulation.
  • capacity (Q) is easily controlled by operating the throttling valve (13). It is further clear from the figure that at stable conditions, the circulating mass flow of CO 2 (m) is mainly constant and independent of the cooling capacity. The CO 2 temperature at the outlet of heat exchanger 11 (t b ) is also mainly constant. The graphs show that the variation of capacity is a result of varying high side pressure (p H ) only.
  • the water inlet temperature to the evaporator is kept constant at 20°C, and the compressor is running at constant speed.
  • the cooling capacity can be kept mainly constant when the ambient temperature is rising, by in ⁇ creasing the high side pressure.
  • the refrigerant mass flow is mainly constant, as shown.
  • Increased high-side pressures involve a reduction in receiver liquid content, as indicated by the liquid level readings.
  • FIG. 7 shows the variation of refrigerating capacity (Q) when the circulation rate of cooling water (m w ) is regulated as shown at the top of the figure.
  • the mass flow of C0 2 (m), the high-side pressure (p H ) and the water inlet temperature to heat exchanger 11 (ti) are kept constant.
  • the compressor is running at constant speed and both the temperature and flow rate of water entering the evaporator are kept constant.
  • the refrigerating capacity is easily controlled by variation of the water flow, as shown in the figure. Mass flow of C0 2 is mainly constant.
  • Fig. 8 is a graphic representation of trans critical cycles in the entropy/temperature diagram. The cycles shown in the diagram are based on measurements on the laboratory test device, during operation at five different high-side pressures. The evaporator pressure is kept constant, refrigerant is C0 2 .
  • the diagram gives a good impression of the capacity control principle, indicating the changes in specific enthalpy (h) at evaporator inlet caused by variation of the high-side pressure (p) .

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical Kinetics & Catalysis (AREA)
  • Chemical & Material Sciences (AREA)
  • Compression-Type Refrigeration Machines With Reversible Cycles (AREA)
  • Devices That Are Associated With Refrigeration Equipment (AREA)
  • Heating, Cooling, Or Curing Plastics Or The Like In General (AREA)
  • Cathode-Ray Tubes And Fluorescent Screens For Display (AREA)
  • Common Detailed Techniques For Electron Tubes Or Discharge Tubes (AREA)
  • Separation, Recovery Or Treatment Of Waste Materials Containing Plastics (AREA)
  • Treatments Of Macromolecular Shaped Articles (AREA)
  • Air-Conditioning For Vehicles (AREA)

Abstract

La présente invention utilise la régulation de l'enthalpie spécifique à l'entrée de l'évaporateur par utilisation délibérée de la pression et/ou de la température avant l'étranglement pour obtenir un réglage de la capacité. La capacité est réglée par variation de la différence d'enthalpie du réfrigérant dans l'évaporateur et par modification de l'enthalpie spécifique du réfrigérant avant l'étranglement. En état surcritique, ce réglage peut être assuré par variation de la pression et de la température séparément. Dans un mode de réalisation préféré, cette modulation de l'enthalpie spécifique est effectuée par variation de la pression avant étranglement. Le réfrigérant est refroidi à une température aussi basse qu'il est possible de descendre au moyen du milieu de refroidissement disponible et la pression est régulée pour produire l'enthalpie requise. Dans un autre mode de réalisation, on utilise la modulation de l'enthalpie par variation de la température du réfrigérant avant étranglement, ce qui est effectué par réglage du rejet de chaleur depuis le dispositif.The present invention utilizes the control of the specific enthalpy at the inlet of the evaporator by deliberate use of pressure and / or temperature before throttling to achieve capacity control. The capacity is regulated by varying the enthalpy difference of the refrigerant in the evaporator and by changing the specific enthalpy of the refrigerant before throttling. In a supercritical state, this adjustment can be ensured by varying the pressure and the temperature separately. In a preferred embodiment, this modulation of the specific enthalpy is effected by varying the pressure before throttling. The refrigerant is cooled to as low a temperature as can be lowered by means of the available cooling medium and the pressure is regulated to produce the required enthalpy. In another embodiment, use is made of modulating the enthalpy by varying the temperature of the refrigerant before throttling, which is done by adjusting the heat rejection from the device.

Description

Title of the invention
Trans-critical vapour compression cycle device
Field of the invention
This invention relates to vapour compression cycle devices such as refrigerators, air-conditioning units and heat pumps, using a refrigerant operating in a closed circuit under trans-critical conditions, and more particularly, to methods for modulating and controlling the capacity of such devices.
Background of the invention
A conventional vapour compression cycle device for refri¬ geration, air-conditioning or heat pump purposes is shown in principle in Fig. 1. The device consists of a compressor (1), a condensing heat exchanger (2), a throttling valve (3) and a evaporating heat exchanger (4). These components are connected in a closed flow circuit, in which a refrigerant is circulated. The operating principle of a vapour compres¬ sion cycle device is as follows: The pressure and tempera¬ ture of the refrigerant vapour is increased by the compres¬ sor (1), before it enters the condenser (2) where it is cooled and condensed, giving off heat to a secondary cool¬ ant. The high-pressure liquid is then throttled to the eva¬ porator pressure and temperature by means of the expansion valve (3). In the evaporator (4), the refrigerant boils and absorbs heat from its surroundings. The vapour at the eva¬ porator outlet is drawn into the compressor, completing the cycle.
Conventional vapour compression cycle devices use refrige¬ rants (as for instance R-12, CF2C12) operating entirely at sub-critical pressures. A number of different substances or mixtures of substances may be used as a refrigerant. The choice of refrigerant is among others influenced by the condensation temperature, as the critical temperature of the fluid sets the upper limit for the condensation to occur. In order to maintain a reasonable efficiency, it is normally desirable to use a refrigerant with critical temperature at least 20-30K above the condensation temperature. Near- critical temperatures are normally avoided in design and operation of conventional systems.
The present technology is treated in full detail in the literature, e.g. the Handbooks of American Society of Heat¬ ing, Refrigerating and Air Conditioning Engineers Inc., Fundamentals 1989 and Refrigeration 1986.
The ozone-depleting effect of todays common refrigerants (halocarbons) has resulted in strong international action to reduce or prohibit the use of these fluids. Consequently there is a urgent need for finding alternatives to the pre¬ sent technology.
Capacity control of the conventional vapour compression cycle device is achieved mainly by regulating the mass flow of refrigerant passing through the evaporator. This is done e.g. by controlling the compressor capacity, throttling or bypassing. These methods involve more complicated flow cir¬ cuit and components, need for additional equipment and ac- cesories, reduced part-load efficiency and other compli¬ cations.
A common type of liquid regulation device is a thermostatic expansion valve, which is controlled by the superheat at the evaporator outlet. Proper valve operation under varying operating conditions is achieved by using a considerable part of the evaporator to superheat the refrigerant, resul¬ ting in a low heat transfer coefficient.
Furthermore, heat rejection in the condenser of the conven¬ tional vapour compression cycle takes place mainly at con¬ stant temperature. Therefore, thermodynamic losses occur due to large temperature differences when giving off heat to a secondary coolant with large temperature increase, as in heat pump applications or when the available secondary cool¬ ant flow is small.
The operation of a vapour compression cycle under trans- critical conditions has been formerly practiced to some extent. Up to the time when the halocarbons took over - 40- 50 years ago - C02 was commonly used as a refrigerant, notably in ships refrigeration for provisions and cargo. The systems were designed to operate normally at sub-critical pressures, with evaporation and condensation. Occasionally, typically when a ship was passing tropical areas the cooling sea water temperature could be too high to effect normal condensation, and the plant would operate with supercritical conditions on the high-side. (Critical temperature for C02 -31°C). In this situation it was practiced to increase the refrigerant charge on the high-side to a point where the pressure at the compressor discharge was raised to 90-100 bar, in order to maintain the cooling capacity at a reason¬ able level. C02 refrigeration technology is described in older literature, e.g. P. Ostertag "Kalteprozesse", Springer 1933 or H.J. Maclntire "Refrigeration Engineering", Wiley 1937.
The usual practice in older C02-systems was to add the ne¬ cessary extra charge from separate storage cylinders. A receiver installed after the condenser in the normal way will not be able to provide the functions intended by the present invention.
Another possibility to increase the capacity and efficiency of a given vapour compression cycle device operating with supercritical high-side pressure is known from German patent 278095 (1912). This method involves two-stage compression with intercooling in the supercritical region". Compared to the standard system, this involves installation of an ad¬ ditional compressor or pump, and a heat exchanger. The textbook "Principles of Refigeration" of W.B Gosney (Cambridge Univ. Press 1982) points at some of the pe¬ culiarities of near-critical pressure operation. It is suggested that increasing the refrigerant charge in the high-pressure side could be accomplished by temporarily shutting the expansion valve, so as to transfer some charge from the evaporator. But it is emphasized that this would leave the evaporator short of liquid, causing reduced capa¬ city at the time when it is most wanted.
Objects of the invention
It is therefore an object of the present invention to pro¬ vide a new, improved, simple and effective means for modu¬ lating and controlling the capacity of a trans-critical vapour compression cycle device, avoiding the above short¬ comings and disadvantages of the prior art.
Another object of the present invention is to provide a vapour compression cycle avoiding use of CFC refrigerants, and at the same time offering possibility to apply several attractive refrigerants with respect to safety, environ¬ mental hazards and price.
Further object of the present invention is to provide a new method of capacity control, which involves operation at mainy constant refrigerant mass flow rate and simple capa¬ city modulation by valve operation.
Still another object of the present invention is to provide a cycle rejecting heat at gliding temperature, reducing heat-exchange losses in applications where secondary coolant flow is small or when the secondary coolant is to be heated to a relatively high temperature. Summary of the invention
The above and other objects of the present invention are achieved by providing a method operating normally at trans- critical conditions (i.e. super-critical high-side pressure, sub-critical low-side pressure) where the thermodynamic properties in the super-critical state are utilized to con¬ trol the refrigerating and heating capacity of the device.
The present invention involves the regulation of specific enthalpy at evaporator inlet by deliberate use of the pres¬ sure and/or temperature before throttling for capacity con¬ trol. Capacity is controlled by varying the refrigerant enthalpy difference in the evaporator, by changing the spe¬ cific enthalpy of the refrigerant before throttling. In the super-critical state this can be done by varying the pres¬ sure and temperature independently. In a preferred embodi¬ ment this modulation of specific enthalpy is done by varying the pressure before throttling. The refrigerant is cooled down as far as it is feasible by means of the available cooling medium, and the pressure regulated to give the re¬ quired enthalpy. Another embodiment involves modulation of enthalpy by variation of the refrigerant temperature before throttling. This is done by controlling the heat rejection from the device.
Brief description of the drawings
The invention will now be described in more detail, in the following referring to attatched drawings, Fig. 1, 2, 3, 4, 5, 6, 7 and 8, where:
Fig. 1 is a schematic representation of a conventional (sub- critical) vapour compression cycle device.
Fig. 2 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a preferred embodiment of the invention. This embodi¬ ment includes a volume as an integral part of the evaporator system, holding refrigerant in the liquid state.
Fig. 3 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a second embodiment of the invention. This embodiment includes an intermediate pressure receiver connected direct¬ ly into the flow circuit between two valves.
Fig. 4 is a schematic representation of a trans-critical vapour compression cycle device constructed in accordance with a third embodiment of the invention. This embodiment includes a special receiver to hold refrigerant as liquid or in the super-critical state.
Fig. 5 is a graph illustrating the relationship of pressure versus enthalpy of the trans-critical vapour compression cycle device of Fig. 2, 3 or 4, at different operating con¬ ditions.
Fig. 6 is a collection of graphs illustrating the control of refrigerating capacity by the method of pressure control in accordance with the present invention. The results shown are measured in a laboratory demonstration system built accord¬ ing to a preferred embodiment of the invention.
Fig. 7 is a collection of graphs illustrating the control of refrigerating capacity by control of the heat rejection, in accordance with the present invention. The results shown are measured in a laboratory demonstration system build accord¬ ing to a preferred embodiment of the present invention.
Fig. 8 is test results showing the relationship of tempera¬ ture versus entropy of the trans-critical vapour compression cycle device of Fig. 2, operating at different high-side pressures, employing carbon dioxide as refrigerant Detailed description of the invention
A trans-critical vapour compression cycle device according to the present invention includes a refrigerant, of which critical temperature is between the temperature of the heat inlet and the mean temperature of heat submittal, and a closed working fluid circuit where the refrigerant is circu¬ lated.
Suitable working fluids may be by the way of examples: ethy- len (C2H4), diborane (B2H6), carbon dioxide (C02), ethane (C2H6 ) and nitrogen oxide (N20).
The closed working fluid circuit consists of a refrigerant flow loop with an integrated storage segment. Fig. 2 shows a preferred embodiment of the invention where the storage segment is an integral part of the evaporator system. The flow circuit includes a compressor 10 connected in series to a heat exchanger 11, a counterflow heat exchanger 12 and a throttling valve 13. The throttling valve can be replaced by an optional expansion device. An evaporating heat exchanger 14, a liquid separator/receiver 16 and the low-pressure side of the counterflow heat exchanger 12 are connected in flow communication intermediate the throttling valve 13 and the inlet 19 of the compressor 10. The liquid receiver 16 is connected to the evaporator outlet 15, and the gas phase outlet of the receiver 16 is connected to the counterflow heat exchanger 12.
The counterflow heat exchanger 12 is not absolutely neces¬ sary for the functioning of the device but improves its efficiency, in particular its rate of response to a capacity increase requirement. It also serves to return oil to the compressor. For this purpose a liquid phase line from the receiver (16) (shown with broken line in Fig. 2) is connect¬ ed to the suction line either before the counterflow heat exchanger (12) at 17 or after it at 18, or anywhere between these points. The liquid flow, i.e. refrigerant and oil, is controlled by a suitable conventional liquid flow restrict¬ ing device (not shown in the figure). By allowing some ex- cess liquid refrigerant to enter the vapour line, a liquid surplus at the evaporator outlet is obtained.
In a second embodiment of the invention indicated in Fig. 3, the storage segment of the working fluid circuit includes a receiver 22 integrated in the flow circuit between a valve 21 and the throttling valve 13. The other components 10-14 of the flow circuit is identical to the components of the previous embodiment, although the heat exchanger 12 can be omitted without any great consequence. The pressure in the receiver 22 is kept intermediate the high-side and low-side pressures of the flow circuit.
In a third embodiment of the invention indicated in Fig. 4, the storage segment of the working fluid circuit includes a special receiver 25, where the pressure is kept between the high-side pressure and the low-side pressure of the flow circuit. The storage segment further consists of the valves 23 and 24 which are connected to the high pressure and low pressure part of the flow circuit respectively.
In operation, the refrigerant is compressed to a suitable supercritical pressure in the compressor 10, the compressor outlet 20 is shown as state "a" in Fig. 5. The refrigerant is circulated through the heat exchanger 11 where it is cooled to state "b", giving off heat to a suitable cooling agent, e.g. cooling air or water. If desired, the refrig¬ erant can be further cooled to state "c" in the counterflow heat exchanger 12, before throttling to state "d". By the pressure reduction in the throttling valve 13, a two-phase gas/liquid mixture is formed, shown as state "d" in Fig. 3. The refrigerant absorbs heat in the evaporator 14 by eva¬ poration of the liquid phase. From state "e" at the evapora¬ tor outlet, the refrigerant vapour can be superheated in the counterflow heat exchanger 12 to state "f" before it enters the compressor inlet 19, making the cycle complete. In the preferred embodiment of the invention, as shown in Fig. 2 , the evaporator outlet condition "e" will be in the two-phase region due to the liquid surplus at the evaporator outlet. Modulation of the trans-critical cycle device capacity is accomplished by varying the refrigerant state at the eva¬ porator inlet, i.e. point "d" in Fig. 5. The refrigerating capacity per unit of refrigerant mass flow corresponds to the enthalpy difference between state "d" and state "e"; This enthalpy difference is found as a horizontal distance in the enthalpy-pressure diagram. Fig. 5.
Throttling is a constant enthalpy process, thus the enthalpy in point "d" is equal to the enthalpy in point "c". In con¬ sequence, the ref igerating capacity (in k ) at constant refrigerant mass flow can be controlled by varying the en¬ thalpy at point "c".
It should be noted that in the trans-critical cycle the high-pressure single-phase refrigerant vapour is not conden¬ sed but reduced in temperature in the heat exchanger 11. The terminal temperature of the refrigerant in the heat ex¬ changer (point "b" ) will be some degrees above the entering cooling air or water temperature, if counterflow is used. The high-pressure vapour can then be cooled a few degrees lower, to point "c", in the counterflow heat exchanger 12. The result is, however, that at constant cooling air or water inlet temperature, the temperature at point "c" will be mainly constant, independent of the pressure level in the high side.
Therefore, modulation of device capacity is accomplished by varying the pressure in the highside, while the temperature in point "c" is mainly constant. The curvature of the iso¬ therms near the critical point result in a variation of enthalpy with pressure, as shown in Fig. 5. The figure shows a reference cycle (a-b-c-d-e-f) , a cycle with reduced capa¬ city due to reduced high side pressure (a'-b'-c'-d'-e-f) and a cycle with increased capacity due to higher pressure in the high side (a"-b"-c"-d"-e-f) . The evaporator pressure is assumed to be constant.
The pressure in the high-pressure side is independent of temperature, because it is filled with a single phase fluid. To vary the pressure it is necessary to vary the mass of refrigerant in the high side, i.e. to add or remove some of the instant refrigerant charge in the high side. These vari¬ ations must be taken up by a buffer, to avoid liquid over¬ flow or drying up of the evaporator.
In the preferred embodiment of the invention indicated in Fig. 2, the refrigerant mass in the high side can be in¬ creased by temporarily reducing the opening of the throt¬ tling valve 13. Due to the incidentally reduced refrigerant flow to the evaporator, the excess liquid fraction at the evaporator outlet (15) will be reduced. The liquid refri¬ gerant flow from the receiver 16 into the suction line is however constant. Consequently, the balance between the liquid flow entering and leaving the receiver 16 is shifted, resulting in a net reduction in receiver liquid content and a corresponding accumulation of refrigerant in the high pressure side of the flow circuit.
The increase in high side charge involves increasing pres¬ sure and thereby higher refrigerating capacity. This mass transfer from the low-pressure to the high-pressure side of the circuit will continue until balance between refrigerat¬ ing capacity and load is found.
Opening of the throttling valve 13 will increase the excess liquid fraction at the evaporator outlet 15, because the evaporated amount of refrigerant is mainly constant. The difference between this liquid flow entering the receiver and the liquid flow from the receiver into the suction line, will accumulate. The result is a net transport of refri¬ gerant charge from the high side to the low side of the flow circuit, with the reduction in the high side charge stored in liquid state in the receiver. By reducing the high-side charge and thereby pressure, the capacity of the device is reduced, until balance is found.
Some liquid transport from the receiver into the compressor suction line is also needed to avoid lubricant accumulation in the liquid phase of the receiver. In the second embodiment of the invention indicated in Fig. 3, the refrigerant mass in the high side can be increased by simultaneously shutting the valve 21 and modulating the throttling valve 13 to provide the evaporator with suf¬ ficient liquid flow. This will reduce the refrigerant flow from the high side into the receiver through valve 21, while refrigerant mass is transferred from the low side to the high side by the compressor.
Reduction of high-side charge is obtained by opening the valve 21 while keeping the flow through the throttling valve 13 mainly constant. This will transfer mass from the high- side of the flow circuit to the receiver 22.
In a third embodiment of the invention indicated in Fig. 4, the refrigerant mass in the high side can be increased by opening the valve 24 and simultaneously reducing the flow through the throttling valve 13. By this, refrigerant charge is accumulated in the high-pressure side due to reduced flow through the throttling valve 13. Sufficient liquid flow to the evaporator is obtained by opening the valve 24.
A reduction in the high side charge can be accomplished by opening the valve 23 to transfer some refrigerant charge from the high side to the receiver. Capacity control of the device is thus accomplished by modulation of the valves 23 and 24, and simultaneously operating the throttling valve 13.
The preferred embodiment of the invention, as indicated in fig. 2 has the advantage of simplicity, with capacity con¬ trol by operation of one valve only. Furthermore, the trans- critical vapour compression cycle device built according to this embodiment has a certain self-regulating capability by adapting to changes in cooling load through changes in liquid content in the receiver 16, involving changes in highside charge and thus cooling capacity. In addition, the operation with liquid surplus at evaporator outlet gives favourable heat transfer characteristics. The second embodiment, as indicated in Fig. 3, has the ad¬ vantage of simplified valve operation. Valve 21 only regu¬ lates the pressure in the high side of the device, and the throttling valve 13 only assures that the evaporator is fed sufficiently. A conventional thermostatic valve can thus be used for throttling. Oil return to the compressor is easily achieved by allowing the refrigerant to flow through the receiver. This embodiment however does not offer the capa¬ city control function at high-side pressures below the cri¬ tical pressure. The volume of the receiver 22 must be rela¬ tively large since it is only operating between the dis¬ charge pressure and the liquid line pressure.
Still another embodiment as indicated in Fig. 4, has the advantage of operating as a conventional vapour compression cycle device, when it is running at stable conditions. The valves 23 and 24, connecting the receiver 25 to the flow circuit, are activated only during capacity control. This embodiment requires use of three different valves during periods of capacity change.
The latter embodiments has the disadvantage of higher pres¬ sure in the receiver, as compared to the preferred embodi¬ ment. The differences between the individual systems regard¬ ing design and operational characteristics are however not very significant.
Trans-critical vapour compression cycle devices built ac¬ cording to the described embodiments can be applied in seve¬ ral areas. The technology is well suitable in small and medium-sized stationary and mobile air-conditioning units, small and medium-sized refrigerators/freezers and in smaller heat pump units. One of the most promising applications is in automotive air-conditioning, where the present need for a new, non-CFC, lightweight and efficient alternative to R12- systems is urgent.
The above described embodiments of this invention are in¬ tended to be exemplative only and not limiting. It will be appreciated that it is also possible to control the capacity of the trans-critical cycle device by keeping the high-side pressure mainly constant, and regulate the refrigerant tem¬ perature before throttling (state "c") by varying the circu¬ lation rate of cooling air or water. By reducing the flow of cooling fluid, i.e. air or water, the temperature before throttling will increase and the capacity will drop. In¬ creased cooling fluid flow will reduce the temperature be¬ fore throttling, and thereby increase the capacity of the device. Combinations of pressure and temperature control are also possible.
Examples
The practical use of the present invention for refrigeration or heat pump purposes is illustrated by the following ex¬ amples, giving test results from a trans-critical vapour compression cycle device, built according to the embodiment of the invention shown in Fig. 2, employing carbon dioxide (C02) as refrigerant.
The laboratory test device uses water as heat source, i.e. the water is refrigerated by heat exchange with boiling CO2 in the evaporator 14. Water is also used as cooling agent, being heated by C02 in the heat exchanger 11. The test de¬ vice includes a 61 ccm reciprocating compressor (10) and a receiver (16) with total volume of 4 liters. The system also includes a counterflow heat exchanger (12) and liquid line connection from the receiver to point 17, as indicated in Fig. 2. The throttling valve 13 is operated manually.
Example 1
This example shows how control of refrigerating capacity is obtained by varying the position of the throttling valve 13, thereby varying the pressure in the high-side of the flow circuit. By variation of high-side pressure, the specific refrigerant enthalpy at the evaporator inlet is controlled, resulting in modulation of refrigerating capacity at constant mass flow.
The water inlet temperature to the evaporator 14 is kept constant at 20°C, and the water inlet temperature to the heat exchanger 11 is kept constant at 35°C. Water circu¬ lation is constant both in the evaporator 14 and the heat exchanger 11. The compressor is running at constant speed.
Fig. 6 shows the variation of refrigerating capacity (Q), compressor shaft work (W), highεide pressure (pH), C02 mass flow (m), C02 temperature at evaporator outlet (t#), C02 temperature at the outlet of heat exchanger 11 (tb) and liquid level in the receiver (h) when the throttling valve 13 is operated as indicated at the top of the figure. The adjustment of throttling valve position is the only mani¬ pulation.
As shown in the figure, capacity (Q) is easily controlled by operating the throttling valve (13). It is further clear from the figure that at stable conditions, the circulating mass flow of CO2 (m) is mainly constant and independent of the cooling capacity. The CO2 temperature at the outlet of heat exchanger 11 (tb) is also mainly constant. The graphs show that the variation of capacity is a result of varying high side pressure (pH) only.
It can also be seen from the diagram that increased highεide pressure involves a reduction in the receiver liquid level (h), due to the C02 charge transfer to the highpressure side of the circuit.
Finally, it can be noted that the transient period during capacity increase is not involving any significant superheating at the evaporator outlet, i.e. only small fluctuations in tβ. Example 2
With higher water inlet temperature to heat exchanger 11 (e.g. higher ambient temperature), it is necessary to in¬ crease the high side pressure to maintain a constant refri¬ gerating capacity. Table 1 shows results from tests run at different water inlet temperature to heat exchanger 11 (tw).
The water inlet temperature to the evaporator is kept constant at 20°C, and the compressor is running at constant speed.
As the table shows, the cooling capacity can be kept mainly constant when the ambient temperature is rising, by in¬ creasing the high side pressure. The refrigerant mass flow is mainly constant, as shown. Increased high-side pressures involve a reduction in receiver liquid content, as indicated by the liquid level readings.
Table 1
Example 3
This example illustrates the possibility to modulate and control the capacity of the device by adjustment of the flow of coolant (e.g. air or water) circulating through heat exchanger 11, keeping the high-side pressure constant. Fig. 7 shows the variation of refrigerating capacity (Q) when the circulation rate of cooling water (mw) is regulated as shown at the top of the figure. The mass flow of C02 (m), the high-side pressure (pH) and the water inlet temperature to heat exchanger 11 (ti) are kept constant. The compressor is running at constant speed and both the temperature and flow rate of water entering the evaporator are kept constant.
The refrigerating capacity is easily controlled by variation of the water flow, as shown in the figure. Mass flow of C02 is mainly constant.
Example 4
Fig. 8 is a graphic representation of trans critical cycles in the entropy/temperature diagram. The cycles shown in the diagram are based on measurements on the laboratory test device, during operation at five different high-side pressures. The evaporator pressure is kept constant, refrigerant is C02.
The diagram gives a good impression of the capacity control principle, indicating the changes in specific enthalpy (h) at evaporator inlet caused by variation of the high-side pressure (p) .

Claims

1. Method for regulation of heating/cooling capacity of a vapour compression cycle device comprising a compressor (10), a heat exchanger (11), a throttling means (13) and a evaporating heat exchanger (14) connected in series forming an integral closed circuit applying a refrigerant operating under trans-critical conditions, c h a r a c t e r i z e d i n t h a t the capacity is regulated and controlled by variation of specific enthalpy of the εupercritically pressurized refrigerant at the inlet of the throttling means (13).
2. The method according to claim 1, c h a r a c t e r i z e d i n t h a t the capacity regulation is performed by variating the supercritical refrigerant pressure at the inlet of the throttling means (13), by variation of the instant refrigerant charge in the high pressure side of the circuit.
3. The method according to claim 1, c h a r a c t e r i z e d i n t h a t the capacity regulation is obtained by variation of the refrigerant temperature at the inlet of the throttling means (13) by controlling the flow rate of the heat exchanging medium absorbing heat in the heat exchanger (11).
4. The method according to claim 2, c h a r a c t e r i z e d i n t h a t the throttling means (13) is applied as steering means to variate the liquid refrigerant inventory of a re¬ ceiver (20) connected between the evaporator (14) and the compressor (10) in the low pressure side of the circuit and that a heat exchanger (12) is included between the receiver (20) and the compressor (10) to exchange heat from the high pressure gas for the pur¬ pose of evaporating liquid supplied from the receiver (20) in order to rapidly increase the charge build-up in the high pressure side without dry-up of the evapo¬ rator (14) and at the same time return oil to the com¬ pressor (.i01 * 6. The method according to claim 2, c h a r a c t e r i z e d i n t h a t variation of instant refrigerant charge in the high pressure side of the flow circuit is obtained by modu¬ lating the valve (21) and the throttling means (13) to vary the supercritically pressurized refrigerant charge in a receiver (22) installed in the flow circuit be¬ tween the valve (21) and the throttling means (13).
7. The method according to claim 2, c h a r a c t e r i z e d i n t h a t" variation of instant refrigerant charge in the high pressure side of the flow circuit is obtained by con¬ tinuously regulating the removal or filling of re¬ frigerant to or from a storage device (25) connected to the high and low pressure sides of the flow circuit by means of pipes with valves (23, 24) and keeping the pressure in the storage device (25) intermediate the high side and the low side pressures.
8. The method according to one or more preceding claims, c h a r a c t e r i z e d i n t h a t the refrigerant is carbon dioxide.
9. The method according to one or more preceding claims, c h a r a c t e r i z e d i n t h a t the trans-critical vapour compresεion cycle device iε applied in automotive air-conditioning.
EP89910211A 1989-01-09 1989-09-06 Method of operating a vapour compression cycle under trans- or supercritical conditions Expired - Lifetime EP0424474B2 (en)

Applications Claiming Priority (3)

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NO890076A NO890076D0 (en) 1989-01-09 1989-01-09 AIR CONDITIONING.
NO890076 1989-01-09
PCT/NO1989/000089 WO1990007683A1 (en) 1989-01-09 1989-09-06 Trans-critical vapour compression cycle device

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WO1990007683A1 (en) 1990-07-12
NO903903D0 (en) 1990-09-07
JPH03503206A (en) 1991-07-18
NO890076D0 (en) 1989-01-09
KR910700437A (en) 1991-03-15
RU2039914C1 (en) 1995-07-20
DE68908181D1 (en) 1993-09-09
DE68908181T4 (en) 1995-06-14
DE68908181T3 (en) 1998-06-18
EP0424474B1 (en) 1993-08-04
KR0126550B1 (en) 1998-04-03
DK214690A (en) 1990-11-06
PL285966A1 (en) 1991-03-25
DK167985B1 (en) 1994-01-10
JPH0718602B2 (en) 1995-03-06
DE68908181T2 (en) 1994-04-14
NO171810B (en) 1993-01-25
EP0424474B2 (en) 1997-11-19
NO903903L (en) 1990-09-07
DK214690D0 (en) 1990-09-07
NO171810C (en) 1993-05-05
UA27758C2 (en) 2000-10-16

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