A HEAT EXCHANGER

The present invention relates to a heat exchanger, and in particular the present invention relates to a heat exchanger for a transcritical refrigeration system.

Transcritical refrigeration systems with CO ₂ as a refrigerant are well known in the art. The critical temperature of CO ₂ is 31.0 ⁰ C and the critical pressure is 73.8 bar. At higher temperatures and pressures no clear distinction can be drawn between liquid and vapour, and CO ₂ is said to be in the so-called super-critical fluid region.

In a conventional refrigeration system, heat release from the refrigerant is based on condensation of the refrigerant. Considering the temperature difference needed in a heat exchanger, i.e. app. 10 ⁰ C, the "real life" upper limit for heat release based on condensation of CO ₂ will be around 20 ⁰ C ambient temperature. Below this temperature, the CO ₂ stays below the critical point and the refrigeration system operates in subcritical cycles.

For refrigeration systems used in supermarkets, the ambient temperature will exceed 20 ⁰ C during the summer in a large part of the world. At these temperatures, cooling of the CO ₂ is a single-phase cooling, namely a gas cooling. CO ₂ is above the critical point at the high- pressure side of the system, and the refrigeration system operates in transcritical cycles.

The efficiency and cooling capacity of the refrigeration system is lower in transcritical operation than in subcritical operation.

It is an important disadvantage of known CO ₂ refrigeration systems that they have a lowered performance at elevated ambient temperatures above app. 20 ⁰ C, i.e. when a high performance is actually desired.

As further explained below, a transcritical refrigeration system with improved performance during transcritical operation is provided by incorporation of one or more heat exchangers.

Typically, the high and low pressure generated by the compressor in a transcritical refrigeration system with CO ₂ refrigerant are approximately 120 bar and 40 bar, respectively.

Thus, there is a need for a heat exchanger that can withstand high pressures and still be manufactured in a simple way at a low cost.

According to the present invention the above-mentioned and other objects are fulfilled by a heat exchanger with a first flow channel positioned in thermal contact with a second flow channel for heat exchange between a first fluid flowing in the first channel and a second fluid flowing in the second channel. The first and second flow channels are tubular flow channels and the first flow channel is positioned within and enclosed by the second flow channel

whereby a good thermal contact between the flow channels is provided facilitating heat exchange between fluids flowing in the flow channels.

Preferably, the tubular flow channels have a circular cross-section for low manufacturing cost. The outer surface of the inner flow channel may be corrugated or have fins so that the surface area is increased and the heat transfer enhanced.

The first tubular flow channel may be substantially concentric with the second tubular flow channel so that the fluid flow in the second channel is substantially uniform and symmetric for the best heat exchange.

In a preferred simple embodiment, the heat exchanger comprises two linear tubes with circular cross-sections extending along the same longitudinal axis.

In another preferred embodiment, the heat exchanger comprises two linear tubes with circular cross-sections extending along the same longitudinal axis that are bend into a meander shape with linear sections interconnected by bended sections, preferably turning the flow direction 180°. In a preferred embodiment, a bended section is formed without thermal contact between the fluid channels so that no heat exchange takes place in the bended section for ease of manufacturing.

In yet another preferred embodiment, the heat exchanger comprises two linear tubes with circular cross-sections extending along the same longitudinal axis that are bend into a spiral.

Preferably, the first channel is formed by a tube made of stainless steel capable of sustaining the pressure of the high-pressure side of the compressor, typically 120 bar. The second channel is preferably formed by a tube made of copper that is silver soldered around the stainless steel tube and having copper fittings forming the flow inlet and outlet. The copper tube is capable of sustaining the pressure of the low-pressure side of the compressor, typically 40 bar. In a preferred embodiment of the system, a heat exchanger according to the present invention is utilized in a transcritical refrigeration system comprising a refrigerant flow circuit for recirculation of a refrigerant, the refrigerant flow circuit comprising a compressor for generation of a refrigerant flow from a low-pressure side to a high-pressure side of the compressor and, in the order defined by the flow direction, connected in series with a gas cooler for cooling of the refrigerant towards the ambient temperature, a pressure reducing device, such as a reduction valve, separating the low-pressure side and the high pressure side of the compressor, and a first evaporator for evaporation of the refrigerant, e.g. in a cooling furniture.

In order to improve the performance of the transcritical refrigeration system during transcritical operation, the heat exchanger according to the invention has its first channel connected in the refrigerant flow circuit between the gas cooler and the pressure-reducing device so that the first channel is incorporated in the refrigerant flow channel. The second channel of the heat exchanger is used for a thermal medium flow. The second channel is connected in series with a storage tank for storage of the thermal medium, e.g. water, the storage tank further comprising a second evaporator connected with the refrigerant flow circuit in parallel with the first evaporator for cooling of the thermal medium.

It is an important advantage of the present invention, that the cooling capacity of the system that is available at low ambient temperatures, i.e. during subcritical operation of the system, may be stored in the storage tank for later use, such as for use during transcritical operation of the system, when the system without the stored cooling capability would operate with a lowered performance. This makes it possible to lower the performance requirements of the compressor in the system, and this lowers the manufacturing cost. Preferably, the thermal medium is water, and preferably the water is cooled to ice in the storage tank during subcritical operation of the system.

Alternatively, the thermal storage tank may also comprise a thermal medium flow path connected in series with the thermal medium flow path of the heat exchanger so that the thermal medium flows into thermal contact with a second thermal medium, e.g. water, stored in the thermal storage tank and cooled by the second evaporator during subcritical operation of the system. The second thermal medium may undergo a phase transition, e.g. water to ice, during cooling.

A phase transition, e.g. from water to ice, makes it possible to store a large cooling capacity in the storage tank. Preferably, the refrigerant is CO ₂ due to its low global warming potential (GWP = 1), availability, and reasonable cost.

For improved performance of the system, a second pressure reducing device may be connected in series with and between the gas cooler and the heat exchanger for adjustment of a desired pressure in the gas cooler. The second pressure reducing device may alternatively be connected between the heat exchanger and the receiver. As will be further explained below with reference to the drawing, an optimum gas cooler pressure exists during transcritical operation of the system. The second pressure reducing device, such as an expansion valve, may be controlled so that the gas cooler pressure attains the optimum value or approximately the optimum value.

The heat exchanger also causes a pressure reduction from the input to the output of the heat exchanger as a result of the cooling of the refrigerant. Thus, the heat exchanger may also function as the second pressure reducing device and the flow of thermal medium through the heat exchanger may be controlled so that the gas cooler pressure attains the optimum value or approximately the optimum value.

A heat exchanger according to the present invention may have the first flow channel connected with the refrigerant flow circuit between the compressor and the gas cooler and with the second flow channel connected with the above-mentioned storage tank for fluid flow of the thermal medium so that heat exchange between the refrigerant at the high pressure side of the compressor and the thermal medium can take place for heating of the thermal medium and storage of the heated medium in the thermal storage tank. The heated medium may be used for heating purposes.

Alternatively, the thermal storage tank may also comprise a thermal medium flow path connected in series with the thermal medium flow path of the second heat exchanger so that the thermal medium flows into thermal contact with a second thermal medium, e.g. water, stored in the thermal storage tank.

A heat exchanger according to the present invention may be used for heat exchange between the high-pressure side and the low-pressure side of the compressor for improved efficiency of the refrigeration system. The first flow channel may be connected in the refrigerant flow circuit between the compressor and the gas cooler, or between the gas cooler and the second pressure reducing device, or between the second pressure reducing device and the receiver. The second flow channel may be connected in the refrigerant flow circuit between the evaporator and the compressor.

Below the invention will be described in more detail with reference to the exemplary embodiments illustrated in the drawing, wherein

Fig. 1 is a blocked schematic of a first embodiment of a transcritical cooling system according to the present invention,

Fig. 2 is a plot of a subcritical cooling cycle, Fig. 3 is a plot of a transcritical cooling cycle, Fig. 4 is a plot illustrating control of gas cooler pressure,

Fig. 5 is a blocked schematic of a second embodiment of a transcritical cooling system according to the present invention,

Fig. 6 shows a longitudinal cross-section of a first and a second embodiment of a heat exchanger according to the present invention, and

Fig. 7 illustrates various positions of a heat exchanger in a refrigeration system according to the present invention.

Fig. 1 is a blocked schematic of a first embodiment 10 of a transcritical cooling system according to the present invention. The system 10 comprises a refrigerant flow circuit for recirculation of CO ₂ refrigerant 12, the refrigerant flow circuit comprising a compressor 14 for generation of a refrigerant flow in the direction of the arrow 16 from a low-pressure side to a high-pressure side of the compressor 14 and, in the order defined by the flow direction, connected in series with a gas cooler 18 for cooling of the refrigerant 12 towards the ambient temperature, a valve 20 for pressure reduction as will be further explained below, a heat exchanger 22 with the first flow channel 24 connected in the refrigerant flow circuit between the gas cooler 18 and a receiver 26 for accommodation of CO ₂ refrigerant 12. The receiver 26 is connected to an expansion valve 28 separating the low-pressure side and the high- pressure side of the compressor 14, and a first evaporator 30 for evaporation of the CO ₂ refrigerant. Fig. 2 illustrates subcritical operation of the system 10 in a conventional Log (p), h (enthalpy) diagram. The enthalpy H is defined by the equation: H = U + pV, where U is the internal energy, p is the pressure, and V is the volume of the system. Between point 1 and 2, the compressor 14 compresses the CO ₂ refrigerant, and subsequently heat is released from the refrigerant from point 2 to 3 below the critical point 32 by condensation of the refrigerant in the gas cooler (condenser) 18 at a constant pressure. The expansion from point 3 to 4 takes place at constant specific enthalpy at passage of the expansion valve 28. The heat absorption takes place in the evaporator 30 in the cooling furniture of the system 10 from point 4 to 1 at constant pressure. The control valve 20 is fully open when the system 10 operates subcritically. Fig. 3 illustrates transcritical operation of the system 10. The most important difference between the plot of Fig. 3 and the plot of Fig. 2 is that the CO ₂ refrigerant is above the critical point 32 at the high-pressure side of the compressor 14 and thus, heat is released from the refrigerant by CO ₂ gas cooling in the gas cooler 18. The coefficient of performance (COP) of the system 10 is less for transcritical cycles than for subcritical cycles due to the lacking phase transition, i.e. no condensation, during heat release.

The expansion from point 3 to 4 takes place in two steps, namely from point 3 to 5, and subsequently from point 5 to 4. The valve 20 reduces the pressure from point 3 to point 5 so that CO ₂ in the liquid phase enters the heat exchanger 22 and is collected in the receiver 26. Further, the valve 20 is controlled in such a way that the pressure in the gas cooler 18 attains a value that gives substantially the best possible COP. This is further illustrated in Fig. 4. In

addition to the transcritical cooling cycle, Fig. 4 shows two isotherms 34, 36. It should be noted that a decrease of the gas cooler pressure at the point 3 moves the point 4 to the right by a large amount because of the low and almost horizontal slope of the isotherm 34 so that the available specific enthalpy for release in the evaporator decrease by a large amount. Since the specific enthalpy added by the compressor 14 decreases by a small amount, the resulting COP decreases by a large amount. Conversely, an increase of the gas cooler pressure at the point 3 moves the point 3 to the left by a small amount because of the steep slope of the isotherm 34 so that the available specific enthalpy for release in the evaporator increases by a small amount. Since the specific enthalpy added by the compressor 14 also increases by a small amount, the resulting COP hardly changes.

It should be noted that if the slope of the isotherm 34 is larger than the slope of the line between points 1 and 2, the COP decreases for increased gas cooler pressure. This illustrates that there is an optimum value for the gas cooler pressure that maximizes the COP, and preferably the valve 20 is adjusted in such a way that the gas cooler pressure attains, at least approximately, this optimum pressure value. Typically, the gas cooler pressure is app. 120 bar while the pressure at the low-pressure side of the compressor 14 is app. 40 bar.

For further improvement of the capacity of the system 10, the heat exchanger 22 also comprises a thermal medium flow channel 38 for a thermal medium flow and connected in series with a storage tank 40 for storage of water and ice. The storage tank 30 further comprises an evaporator 42 connected to an expansion valve 44 so that the evaporator 42 is connected in the refrigerant flow circuit in parallel with the first evaporator 30. During subcritical operation, the evaporator 42 operates in parallel with the evaporator 30 so that the water in the storage tank 20 is cooled to ice. During transcritical operation the pump 46 is operated and the three-way valve 50 is opened to pump water at the freezing point through the thermal medium flow channel 38 for further cooling of the CO ₂ refrigerant whereby the capacity of the system 10 is increased during transcritical operation. The phase transition from water to ice, and vice versa, makes it possible to store a large cooling capacity for later use. The embodiment of Fig. 5 corresponds to the embodiment of Fig. 1 with a further heat exchanger 50 inserted in the refrigerant flow circuit between the compressor 14 and the gas cooler 18. The heat exchanger 50 has a refrigerant flow channel 52 connected in the refrigerant flow circuit between the compressor 14 and the gas cooler 18. The heat exchanger 50 also comprises a thermal medium flow channel 54 for a thermal medium flow and connected in series with the storage tank 40 for storage of water and ice. Thus, in the winter time when the system 10 operates in subcritical cycles, the valves 48, 56, 58 are

opened allowing the pump 46 to pump water from the storage tank 40 through the heat exchanger 50 for heating of the water, and back into the storage tank. The storage tank may further be connected to a heating system (not shown) that utilizes the heated water. During the summer time, the water is cooled to ice, when the system 10 operates in subcritical cycles and the ice water is used for cooling of the refrigerant in the heat exchanger 22 when the system 10 operates in transcritical cycles as described above with reference to Fig. 1.

Fig. 6 shows a longitudinal cross-section of a first and a second embodiment of a heat exchanger according to the present invention.

The heat exchanger 100 has a first flow channel 102 positioned in thermal contact with a second flow channel 104 for heat exchange between a first fluid flowing in the first channel 102 and a second fluid counter-flowing in the second channel 104. The first flow channel is defined within a stainless steel tube 106 with a circular cross-section that is capable of sustaining high pressures, such as the pressure of the high-pressure side of the compressor, e.g. 120 bar. The second channel is defined within a copper tube 108 between the inner wall of the copper tube 108 and the outer wall of the stainless steel tube 106 that is positioned inside the copper tube 108. Heat exchange takes place through the stainless steel wall of the stainless steel tube 108 having good heat conducting properties and a large contact surface with the fluid in the copper tube 108. The inlet 110 and outlet 112 of the copper tube are formed by copper fittings. The stainless steel tube 106 and the copper tube 108 are silver soldered together at the ends 114, 116 of the copper tube and extend concentrically along a common centre axis.

In another preferred embodiment, illustrated in the lower part of Fig. 6, the heat exchanger 100 comprises two linear tubes 106, 108 with circular cross-sections extending along the same longitudinal axis that are bend into a meander-like shape with two linear sections 118, 120 interconnected by a bended section 122 turning the flow direction 180°. The heat exchanger may have more than two linear sections with a corresponding number of bended sections.

Fig. 7 illustrates various positions of a heat exchanger in a refrigeration system according to the present invention. In position 1 , a heat exchanger according to the present invention is used for heat exchange between the high-pressure side and the low-pressure side of the compressor for improved efficiency of the refrigeration system. The first flow channel is connected in the refrigerant flow circuit between the gas cooler and the second pressure reducing device. The second flow channel is connected in the refrigerant flow circuit between the evaporator and the compressor.

Position 2 corresponds to the position of the heat exchanger in Fig. 1. The heat exchanger may also be positioned upstream the second pressure reducing device. Position 3 corresponds to the position of the second heat exchanger in Fig. 5.

An embodiment of the invention may have one heat exchanger in any of the positions shown in Fig. 7 or upstream the second pressure reducing device as required. Another embodiment may have two heat exchangers in any combination of the positions in Fig. 7 or upstream the second pressure reducing device as required. Yet another embodiment may have three heat exchangers in the positions shown in Fig. 7 or upstream the second pressure reducing device as previously mentioned.