Conventional vapour-compression-cycle devices use refrigerants (as for in- stance R22, R134a, R404a, CF2CI2) operating entirely at sub-critical pres- sures. A number of different substances or mixtures of substances may be used as refrigerant. The choice of refrigerant is among others influenced by condensation temperature, as the critical temperature of the fluid sets the upper limit for the condensation to occur. This technology is treated in full details in the literature, e. g. in the Handbooks of American Society of Heat- ing, Refrigerating and Air Conditioning Engineers Inc., Fundamentals 1989 and Refrigeration 1986.

The operation of a vapour-compression-cycle device under trans-critical conditions have been practiced to some extent. Until the use of halocarbons took over in the fifties, C02 (Carbon dioxide) was a generally known refriger- ant. C02 refrigeration technology is described in earlier literature, e. g. P. Os- tertag"Kaiteprozesse", Springer Verlag 1933 or H. J Macintire"Refrigeration Engineering", Wiley 1937.

Vapour-compression-cycle devices operating under trans-critical conditions are known from German patent 278095 from 1912. This patent describes a method involving two-stage compression of a refrigerant, with intercooling of said refrigerant in the supercritical region. Two-stage compression involving intercooling is ideal when the process employs a cooling agent that operates under trans-critical conditions as the temperature generated under the com- pressing process would reach high levels and influence the performance of the cycle. However, the method described suffers from the drawback that the

components are placed in a manner that does not comply with today's de- mands for high efficiency products/processes.

A further vapour-compression-cycle device operating under trans-critical conditions is known from patent application No. PCT/N089/00089 that de- scribes a vapour-compression-cycle device and wherein the capacity regula- tion of said device employs a single-stage compression cycle process com- prising a single-speed compressing device, where the capacity is regulated by varying the refrigerant-enthalpy difference in the evaporator by changing the specific enthalpy of the refrigerant before throttling. This modulation of specific enthalpy is achieved by varying the pressure in the high-pressure side of the compression cycle device that adjusts the throttling means open- ing.

It is generally known that the commonly used refrigerants (halocarbons), when employed in a vapour-compression-cycle device, usually ensures a high COP (coefficient of performance) for the device in question. The COP (coefficient of performance) is defined as the ratio between the power (kW) consumed for driving the compressing devices and the cooling capacity (kW) produced by the vapour-compression-cycle device. For vapour- compression-cycle devices employing for instance R134a, R4Q4a, a ratio of 1: 3-4 is not uncommon. For vapour-compression-cycle devices employing C02 the COP is approximately 1: 2-3.

Due to the ozone-depleting effects of the commonly used refrigerants (halo- carbons) international actions have been taken to reduce or prohibit the use of these gasses. As a consequence, there is a need to find alternatives to the known technology that are environmentally friendly and still economical in all aspects.

It is therefore an object of the present invention to provide a novel, improved, efficient and simple means for regulating the capacity of a trans-critical va-

pour-compression-cycle device avoiding the above shortcomings; and to im- prove the efficiency of the vapour-compression-cycle process.

Another object of the present invention is to provide a vapour-compression- cycle device employing a refrigerant that operates under trans-critical condi- tions and with the smallest possible power consumption; and to employ at- tractive refrigerants with respect to environmental hazards.

The above and other objects of the present invention are achieved by pro- viding a method operating at usually or usual trans-critical conditions, where the capacity is regulated by regulating the number of revolutions of the drive shaft of the compressing device as a function of the required capacity; and wherein the variable flow control means is regulated as a function of the temperature and/or pressure of the refrigerant measured at the high pres- sure side of the system with respect to optimizing the COP (coefficient of performance) of the system.

Another object of the present invention is achieved by providing regulation means for the vapour-compression-cycle device that comprises means for regulating the revolutions of the drive shaft of said compressing device for regulating the capacity; and means for controlling said variable control means as a function of the temperature of the refrigerant on the high pres- sure side with respect to optimizing the COP of the system.

Still another object of the present invention is to provide a control means, said control means for the conditions given regulating the throttling means on basis of temperature/pressure of the refrigerant on the high-side for opti- mizing the COP (coefficient of performance) Yet a further object of the present invention is to provide a vapour- compression-cycle device comprising two or more compression stages. By introducing two or more compression stages, it has been proven that an ac- curate regulation is possible by utilizing more simple specifications of the algorithm for the optimal pressure after the variable flow control means.

A further object of the present invention is to provide a method for regulating the regulation means for the variable flow control means at decreasing tem- perature of the refrigerant arranged to regulating said variable flow control means until a temperature/pressure inferior to the critical temperature, whereupon the opening for refrigerant of the variable flow control means is fully opened, and at increasing temperatures the regulation means starts to regulate said variable flow control means at a temperature higher than the critical temperature.

By this method it is ensured that it will be possible to control and regulate the variable flow control means at temperatures of the refrigerant/gas about the critical value of ca. 32°C in order to prevent to many switching in/off's of the compressing devices.

A further object of the present invention is to provide a method, wherein the refrigerant, after the first compression stage and before the second com- pression stage, is cooled in a first cooler that employs air as coolant, said first cooler being with respect to the airflow placed in series, and after a sec- ond cooler that employs air as coolant for cooling/condensing the refrigerant after the second stage compression.

Yet a further object of the present invention is to provide a method for regu- lating a transcritical vapour-compression-cycle device employing C02 (car- bon dioxide) as refrigerant.

Yet another object of the present invention is to provide a device comprising means for measuring the temperature of the refrigerant on the high-pressure side with respect to optimizing the COP of the system as a function of the temperature or pressure.

Yet another object of the present invention is to provide a vapour- compression-cycle device comprising a cooler employing water as coolant arranged in the integral closed circuit after a compressing device and before the air-cooled cooler.

By an arrangement with the suggested configuration, it is achieved that the second cooler (E) that employs air as coolant does not transfer heat from the circulation air and on to the first cooler (C) that employs air as coolant. Due to a high temperature of the circulation air, the heat transmission and by this the capacity of the first cooler would then approach a zero effect.

Still another object of the present invention is to provide a vapour- compression-cycle device comprising a receiver arranged in the integral closed circuit after the variable flow control means but before the econo- mizer. With the suggested configuration, it will be achieved that the receiver accumulates the refrigerant in liquid state. The pressure elevated in the re- ceiver will at any time be the pressure generated when the refrigerant ex- pands adiabatically from the gas cooler to the threshold curve.

Another object of the present invention is to provide a vapour-compression- cycle device employing CO2 (carbon dioxide) as refrigerant by which a re- frigerant operating under trans-critical conditions is environmentally friendly.

Yet another object of the present invention is to provide a vapour- compression-cycle device comprising coupling means for establishing water cooling, by which it will be made possible to establish water-cooling when the container is loaded onboard a ship, and air cooling of the heat exchang- ers/condensers is not sufficient for cooling the refrigerant/gas.

A further object of the present invention is to provide a vapour-cycle- compression device comprising means for coupling to a ship's electrical power system, and wherein the compression device is driven by an electrical motor means coupled to said electrical power system, by which it is ensured that the vapour cycle compression device is operational when loaded on a ship for transportation.

Yet another object of the present invention is to provide a method employed for refrigerating the interior of a refrigerated container. By this it has been achieved that goods requiring a refrigerated atmosphere can be trans- ported/stored in a container that provides such refrigerated atmosphere.

Brief description of the drawings The invention will now be described in brief terms in the following with refer- ence to the drawings, Figure where: Figure 1 is a graph illustrating the vapour compression cycles for the refrig- erants R134a and CO2 (carbon dioxide); and Figure 2 is a graph illustrating the capacities (Qo) dependency of the temperature of the refrigerant gas/vapour at output of the cooler at maintained pressure; Figure 3 is a graph illustrating the capacity (Qo) dependency of the pressure in the cooler for maintained gas temperature; Figure 4 illustrates calculated pressure in the cooler for optimal COP de- pending off the evaporation temperature and gas temperature at output of the cooler; Figure 5 illustrates calculated pressure in the cooler for optimal COP de- pending off on evaporation temperature and the gas temperature at the out- put from the cooler; Figure 6 illustrates calculated sensitivity for correctly selected pressure in the cooler for optimal COP (TO =-25 °C) ; Figure 7 is a graph illustrating an H-log P diagram for at two-stage system operating at trans-critical conditions; Figure 8 is a graph illustrating a H-log P diagram for at two-stage system operating at sub critical conditions ;

Figure 9 is a calculated comparison of gas temperature; and Figure 10 is a diagram for determining the adjustment pressure; and Figure 11 is a schematic representation of a trans-critical vapour- compression-cycle device according to a preferred embodiment of the inven- tion.

In the figures letters refer to items described and numbers refer to the vari- ous stages of the refrigerant in an enthalpy/temperature diagram, in the fol- lowing figures.

When C02 is used as refrigerant in vapour-compression-cycle device or heat pumps, the process will depend on the temperature at the heat ex- changer/condensing unit being above or below the critical temperature. At temperatures of the coolant in the condenser/heat exchanger below the criti- cal value, the integral circuit process of the refrigerant (CO2) will not be dif- ferent from the integral circuit process of other refrigerants complying to the Carnot cycle/process. However, at temperatures higher than the critical value, the integral circuit process will be different as C02 cannot condense at temperatures above 31 °C. This does not mean the process cannot supply cooling or heating, only that the process is to be configured to comply with another integral circuit process, namely the Lorentz proces. The condenser will no longer be used for condensing the refrigerant, but solely for cooling of the trans critical coolant/fluid, and is hence described as the"gas-cooler" In figure 1 the trans critical integral circuit process is illustrated in compari- son to the conventional integral circuit process employing R134a as refriger- ant. Listed in a H, log (P) -diagram (H for enthalpy, P for pressure). In the fig- ure both isotherms (40 °C) for both refrigerants-C02 (carbon dioxide) and R134a) her mangler noget. Both processes operate at an evaporating tem- perature at minus 10 °C and up against an outdoor temperature or a water temperature of the coolant of approximately 40 °C. As it will appear from Figure 1, the integral circuit process employing CO2 operates at a pressure lever higher than the integral circuit process employing R134a. This means that pipe systems, receivers, and components in general are to be config- ured for this much higher operating pressure. It will further appear from Fig- ure 1 that the expansion vessel/throttling means have supercritical fluid at the inlet where it normally for R134a will be a liquid. This means that the liq-

uid formation for the trans critical process is taking place in the expansion vessel/throttling means during the expansion though the nozzle, which nor- mally do not give reason for problems At operation at trans-critical operations, there is no connection (dependency) between pressure and temperature. This means that the system has been awarded yet another degree of freedom, in the sense that it will be possible to control/regulate the pressure in the gas-cooler and the temperature inde- pendently.

Figure 2 illustrates how the cooling capacity is increased dramatically by changing the temperature of the refrigerant (CO2) by discharge from the gas cooler. The work of the compressing device (Wk) remains the same. Under normal conditions the temperature of the C02 at output of the gas-cooler cannot be chosen arbitrarily, but will be dependent of the conditions of the integral circuit process device given. This may be the temperature of the air, when air is employed as cooling agent, or the temperature of the cooling wa- ter, when water is employed as cooling agent. Since it is not possible to con- trol the temperature of the refrigerant at output of the gas cooler, this means that instead a pressure optimal for a requested run (low/high capacity) or just for the conditions given, to maximize COP (coefficient of performance) for the integral circuit process device. In some cases it may be relevant to re- duce the flow of coolant (air/water) over the gas cooler/condenser in order to reduce the capacity. The result may then be that the COP (coefficient of per- formance) of the integral circuit process device is not optimal ; however, these steps will prevent many start/stops of the integral circuit process de- vice.

Capacity regulating of a trans-critical integral circuit process device employ- ing C02 as refrigerant can be achieved-apart from the normal methods-by controlling the pressure in the gas-cooler. For a temperature of the C02-re- frigerant given at output of the gas cooler, (determent by exterior conditions)

it is thus possible to regulate the capacity of the integral circuit process de- vice within a wide range. The principle is illustrated by Figure 4.

Assuming the temperature of the refrigerant (CO2) at discharge of the gas cooler/heat exchanger is determent by the exterior conditions and has an value of 35 °C as will appear from Figure 3, it will then be possible to in- crease the capacity of the integral circuit process device considerably by raising the pressure in the gas cooler/heat exchanger slightly (en smule).

The other way around, it will be possible to reduce the capacity of the inte- gral circuit process device should this be required, where at the same time the energy consumption of the compressing device is reduced and a energy retrenchment is obtained.

There is often no need for this type of capacity regulation or it may be achieved by other means, for instance by capacity regulations of the com- pressing device (cylinder switch-off or regulation of number of revolutions) In this situation it will be possible to regulate in respect to the maximal COP of the process.

Figure 4 illustrates the calculated optimal choice-in respect to maximal COP-of the gas pressure in the gas cooler/heat exchanger dependent of the temperature of the refrigerant gas at output from the gas cooler. As it will appear from Figure 4 the optimal high pressure is dependent of both the gas temperature at output of the gas cooler, but even of the evaporation tem- perature. Further the optimal pressure will depend on the mechanical effi- ciencies of the compressing devices and pressure losses in exchangers and pipes. On the background of an analysis, the optimal high-pressure can be found on the basis of varying evaporation temperatures (Te) and tempera- tures at output of the gas cooler (Tg, 2). The isentropic efficiency compressing device is kept at 0,6.

Figure 5 illustrates calculated pressure in the cooler for optimal COP de- pendent of evaporation temperature and the gas temperature at the output from the cooler. On this basis a mathematical formula has been expressed, stating the optimal pressure of the gas/refrigerant at output of the cooler, for optimal COP on the basis of Te and Tg, 2.

Pg, 2 (optimal) = 0,7244-0, 275 * Te + 2,275 * Tg, 2 In figure 6 the sensitivity for correct chosen high-pressure is illustrated. As it will appear from the curves, choosing a operating pressure below the optimal pressure as this result in a low COP. The slope is relatively high at these pressures. It will however appear that for pressures higher than the optimal operating pressure, the slope is low. The optimum of each curve represented by a condensing temperature is connected by a broken line in their vertex.

Figure 7 is a graph illustrating a H-log P diagram for a two-stage system op- erating at trans-critical conditions. The numbers are referring to the stages of the circle process, where: (1) is the condition of the refrigerant at the inlet of the first compressing device (A). (2) the condition after the compression by the low-pressure compressing device (A), the condition of the refrigerant at (3) is after cooling of the gas in the first cooling device (C) and 4 is the con- dition at the inlet of the high-pressure compressing device (B). The pressure is then elevated to the pressure at (5) after compression in the high-pressure compressing device (B). Hereafter the refrigerant is cooled reaching the stage at (6) after cooling/condensation in the gas cooler/condensing unit (E).

In the receiver the pressure of the refrigerant is reduced at the stage with number (7). Hereafter the refrigerant is cooled further by the economizer, by which some sub-cooling is ensured reaching the stage (8), which indicates the condition of the refrigerant before the evaporating in the evaporator.

However, for two-stage vapour compression cycle devices plants compre- hensive calculations have shown that it is possible to obtain sufficient accu-

racy by use of a more simple description of the algorithm for the optimal gas pressure after the variable control means. See Figure 10.

Calculations have been made for the suction pressure on 10,20 and 30 bar, respectively, which is the maximal variation for the suction pressure in a va- pour compression cycle device for transportation purposes. As is the case when one-stage vapour-compression-cycle devices are concerned, calcula- tions show that the optimal depends on the suction pressure, but the de- pendency is less severe than that of one-stage devices. This is due to the fact that the suction pressure for the second compressing stage varies only half as much as for the first stage. Therefore, the deliberations have been whether it is possible to replace the fairly complicated expression with a simpler one and use a curve that is independent of suction pressure. When the curve is used for a 20-bar suction pressure, it has been found that the theoretical mistake made that the curve is used for a 20-bar suction curve only is less than 1 % at a gas temperature of 50°C, decreasing to zero at the critical pressure of 32°C where all the curves collapse. In practice the gas temperature will never exceed 45°C.

It is worth noticing that the variable flow control means continues to be a controlling factor for the gas temperature, decreasing until 27. 5°C, following which it must be completely open. Conversely, at increasing gas temperature a regulation is not initiated until the temperature exceeds 32. 5°C. This func- tion is necessary to avoid too excessive switch in and switch outs at gas temperatures around the critical temperature of about 32°C.

Calculation of the optimal gas pressure as disclosed in Figure 11 have in- cluded the values for isentropic efficiency found by tests on current vapour- compression-cycle devices engaging C02 as refrigerant.

It is a further characteristic that gas pressure and temperature within the re- ceiver will always settle at a level that corresponds to the saturation pres-

sure that prevails where the gas expands, at constant enthalpy, above the throttle valve to upper or lower threshold curve.

Figure 11 shows a schematic representation of a trans-critical vapour- compression-cycle device according to a preferred embodiment of the inven- tion, comprising two-stage compression of a refrigerant. In the integral closed circuit the vapour-compression-cycle device comprises a first com- pressing device (A) with an inlet and an outlet side, connected in series to a first heat exchanger (C) and employing air as coolant; and a subsequent second compressing device (B) with an inlet and outlet side, connected to a counter flow heat exchanger (D) and employing water as coolant. In the closed circuit after said counter flow heat exchanger (D), a second heat ex- changer (E) employing air as coolant is arranged. Said first and second heat exchangers (C, E) employing air as coolant are, with respect to the airflow, arranged in series with the second heat exchanger (E) first and the first heat exchanger (C) second, arranged in a trunk-like or similar space providing a air path for the cooling air. Hereafter a ventilator fan is arranged to ensure a sufficient airflow for the heat exchanging process. In the integral closed cir- cuit after the heat exchanger (E), a sensor device followed by a variable flow control means (F) comprising regulation means (not shown) and a receiver (G) are arranged, followed by an economizer (H) with an inlet and outlet, and a throttling means (expansion vessel) (J) comprising a sensor attached to the suction line, described in the latter. In the integral closed circuit, an economizer by-pas line comprising an inlet and outlet is connected with en- trance after the economizer (H) and before the throttling means (expansion vessel) (J) and the outlet connected to the integral closed circuit after the first heat exchanger (C) and before the second compression device (B). In the economizer by-pass line, a throttling means (expansion vessel) (I) is ar- ranged before the economizer, said throttling means (expansion vessel) (I) with the sensor being attached to the in the economizer by-pass line after the economizer (H). The function of throttling means (expansion vessel) (I) will be known to the skilled person, hence it will not be described in further

detail in the following. it is the function of the economizer (H) to ensure some degree of sub-cooling of the refrigerant. In the closed circuit after the throt- tling means (expansion vessel) (J) an evaporator (K) with an inlet and an outlet is connected to the inlet connected with the throttling means (expan- sion vessel) (J) and the outlet to the side connected to the inlet of the first compression device (A), by a line/pipe hereafter described as suction line.

The function and arrangement of throttling means (expansion vessel) (J) and sensor will be known to the skilled person, hence it will not be described in further detail in the following. It will be understood that the evaporator (K) is installed inside a container in which a refrigerated atmosphere is required as it will be likewise understood that a ventilation fan (L) is advantageously ar- ranged to circulate air around the evaporator (K).

The first compression device (A) and the following second compression de- vice (B) both comprising a drive shaft are, in the representation of Figure 1, shown as a two-stage compressing device unit, sharing the same drive shaft, said drive shaft being coupled to a motor means. It will, however, be under- stood that the above described embodiment of the invention is intended to be exemplary only and not in any way limiting the scope of the invention. It will be appreciated that the two-stage compressing device driven by one mo- tor means can be replaced by two single stage compressing devices each comprising a motor means.

The vapour-compression-cycle device comprises a first sensor means ar- ranged on the high-pressure side in the closed circuit before the variable flow control means (F) and after the second cooler (E). Another sensor means is arranged in the closed circuit after the evaporator and before the first inlet of the compression device (A). The sensor means is, in a preferred embodiment, temperature sensors, monitoring the temperature condition of the refrigerant/gas and giving a input signal to the vapour-compression-cycle devices regulation means for regulating the capacity of the vapour- compression-cycle device.

The vapour-compression-cycle device according to a preferred embodiment of the invention is suitable for installation in a container of the type used on in particular ships, for the transportation of goods that require refrigerated surroundings. When the device is loaded onboard, the ship's power system is connected to provide the drive motor for the compressing devices (A, B) and all other electrical components that are powered electrically, and the counter flow heat exchanger (D) is connected to a cooling water system, said cooling water system providing cold water for containers equipped with a va- pour-compression-cycle device comprising a heat exchanger employing wa- ter as cooling agent/coolant. After connection to said systems, the ventilation fan (N) is turned off and the cooling of the refrigerant after the second com- pression device (B) is taken over from the second heat exchanger (E) by the counter flow heat exchanger (D). An airflow as the result of convexion in the airpath ensures some amount of cooling air or the first heat exchanger (C) The practical problem occurs when the same vapour-cycle-compressing- device is set to handle both super-and sub-critical pressure. The regulation function must be capable of handling this conversion. At sub critical opera- tion, the variable control means, in principle must be fully open and in super critical operation adjustable. It will however be practical possible to make this conversion pass without dramatic variations in the temperature of the re- frigerant.

The regulation of the variable flow control means is illustrated in figure 9, listing the comparison between the gas temperature at output gas cooler and optimal gas pressure, and with the regulating curve sketched, said regulation curve determent by the following equation. a = 282.05394 b =-1. 5741522 c =-167064. 7

d = 425285.17 e =-3. 68426 x 109 <BR> <BR> <BR> <BR> <BR> Preg = a + b # Tg,2 + c # In(Tg,2) + d + e-Tg,2<BR> <BR> <BR> Tg,2 Tg,2 For Tg2 smaller than 20 °C the variable flow control means is kept open. For Tg2<20°C, then Tg2=20°C.