FIELD OF THE INVENTION

The present invention relates to a refrigeration system comprising a compressor for compressing a gaseous refrigerant, such that the temperature and pressure thereof increase, whereas the boiling point thereof decreases, a condenser, in which the gaseous refrigerant from the compressor exchanges heat with a high temperature heat carrier, said heat exchange resulting in the refrigerant condensing, an expansion valve reducing the pressure of liquid refrigerant from the condenser, hence reducing the boiling point of the refrigerant, an evaporator, in which the low boiling point refrigerant exchanges heat with a low temperature heat carrier, such that the refrigerant vaporizes, and a suction gas heat exchanger exchanging heat between high temperature liquid refrigerant from the condenser and high temperature gaseous refrigerant from the evaporator. PRIOR ART

In the refrigeration field, there is a constant strive towards more efficient systems.

Actually, the best refrigeration systems approach the Carnot efficiency, which is the theoretical upper limit for a heat machine. Generally speaking, all refrigeration systems transforming mechanical energy to a temperature difference comprise a compressor, a condenser, an expansion valve, an evaporator, and piping enabling transport of refrigerant between the compressor, the condenser, the expansion valve and the evaporator, wherein heat is transferred from the evaporator to the condenser.

However, although the efficiency at some temperature differences may approach the Carnot efficiency, this is far from true for all running conditions.

In general terms, all heat exchangers comprised in a refrigeration system should be as large and efficient as possible. Also, they should have an as low hold-up volume as possible, and a low pressure drop. As could be understood, these criteria cannot all be met.

When it comes to the temperatures after the evaporator, every temperature increase over the temperature at which all refrigerant is evaporated (i.e. the highest boiling point of the refrigerant) will mean a loss in efficiency - however, since liquid refrigerant entering the compressor may seriously damage the compressor, it is also crucial that all refrigerant is actually vaporized before entering the compressor. A state where the entire refrigerant is evaporated, although its temperature does not exceed the boiling temperature, is generally referred to as "zero superheat", and is a state being very beneficial in terms of efficiency.

One way of achieving "zero superheat" in the evaporator is to "flood" the evaporator with liquid refrigerant and let refrigerant boil off from the flooded evaporator. This configuration is common in large chiller applications, i.e. heat machines having a power of 500-1000 kW. Usually, so-called "plate and shell" or "shell and tube" heat exchangers are used for such applications.

As could be understood from the above, such evaporator configurations give great performance, but they are far from free from drawbacks. Firstly, all heat exchangers comprising a shell are bulky and heavy, meaning that the material cost for manufacturing them are high. Secondly, and even more importantly, the refrigerant volume required for flooding the heat exchanger is large. In addition to the cost issue, legislation often bans too large refrigerant amounts in a heat machine.

The by far most efficient heat exchanger type in terms of heat transfer/material mass is the compact brazed plate heat exchanger (BPHE). As known by persons skilled in the art, such heat exchangers comprise a number of plates made from sheet metal and provided with a pressed pattern of ridges and grooves adapted to keep the plates on a distance from one another under formation of interplate flow channels for the media to exchange heat. The plates are brazed to one another, meaning that each plate pair will be active in containing the refrigerant under pressure in the heat exchanger. Brazed plate heat exchangers have the benefit that virtually all material in the heat exchanger is actually active for heat exchange, unlike the heat exchangers comprising a shell, wherein the shell has the sole purpose of containing the refrigerant.

The evaporation processes in BPHE:s and flooded shell and tube heat exchangers are very different - as mentioned, the evaporation in a flooded shell and tube heat exchanger resembles a pool boiling, whereas in a BPHE, the refrigerant will travel more or less linearly through the interplate flow channel. The closer to the exit, the less liquid refrigerant will be present. Because of the volumetric increase due to evaporation, the velocity and hence flow resistance will increase along the length of the heat exchanger.

As mentioned above, it is crucial that no liquid refrigerant enters the compressor. Therefore, it is not uncommon that at least some of the heat exchanger contains only gaseous refrigerant. The gaseous refrigerant will take up heat and become unnecessarily hot, which will decrease the system efficiency.

It is also beneficial if the liquid refrigerant about to enter the evaporator is cool, since flash boiling phenomena can be minimized if the refrigerant is cool.

One way of securing a low refrigerant temperature of the refrigerant about to enter the expansion valve (hence reducing the risk of flash boiling), while securing a high enough temperature of the gaseous refrigerant about to enter the compressor, is to use a so-called suction gas heat exchanger. In its simplest form, a suction gas heat exchanger may be arranged by simply placing the piping from the evaporator to the compressor in the vicinity of the piping from the condenser to the expansion valve close to one another and braze or solder them together, such that heat may be transferred between the pipings. For larger systems, however, it is more common to provide a more efficient heat exchanger than simply two pipes placed beside one another.

If the superheating of the refrigerant could be kept at a minimum while it is ensured that no liquid refrigerant enters the compressor, the BPHE could be competitive with the flooded shell and tube heat exchanger also in terms of efficiency, while retaining its benefits in terms of compactness and material efficiency.

It is the object of the present invention to provide a system where a BPHE is used in a system offering zero, or close to zero, superheat of refrigerant entering the compressor.

SUMMARY OF THE INVENTION

The above problems with the prior art refrigeration systems are solved, or at least alleviated, by a refrigeration system according to the preamble of the attached independent claim, further comprising means for controlling the amount of heat exchange in the suction gas heat exchanger. According to one embodiment, the means for controlling the amount of heat exchange in the suction gas heat exchanger may a controllable by -pass valve, which controls the amount of refrigerant bypassing the suction gas heat exchanger. This embodiment is beneficial in that the refrigeration will be relatively easy to control.

In order to enable use of small diameter valves, the bypass valve may bypass liquid refrigerant from the condenser.

In another embodiment, the means for controlling the amount of heat exchange in the suction gas heat exchanger comprises dual expansion valves, wherein a first of the expansion valves is connected between an inlet of the evaporator and the suction gas heat exchanger and a second of the expansion valves is connected between the inlet of the evaporator and the condenser. This embodiment is beneficial in that it allows full bypass of the suction gas heat exchanger. The dual expansion valves may be controllable. BRIEF DESCRIPTION OF THE DRAWINGS

In the following, embodiments of the invention will be described in detail with reference to the appended drawings, wherein

Fig. 1 is a schematic plan view showing a refrigeration system according to a first embodiment;

Fig. 2 is a schematic plan view showing refrigeration systems according to a second and a third embodiment; and

Fig. 3 is a schematic plan view showing a refrigeration system according to a fourth embodiment. DESCRIPTON OF EMBODIMENTS

In Fig. 1, a refrigeration system according to a first embodiment is shown. The refrigeration system according to the first embodiment comprises a compressor C, an outlet of which being connected to an inlet opening CI of a condenser COND. This inlet opening is in fluid communication with an outlet opening C3. An inlet opening C4 is in fluid communication with an outlet opening C2, and a heat carrier flows between these inlet and outlet openings under heat exchange with initially gaseous refrigerant delivered from the outlet of the compressor C to the inlet CI of the condenser COND. Due to the heat exchange, the initially gaseous refrigerant will condensate during its passage between the inlet CI and the outlet C3.

The outlet C3 is fluidly connected to both an inlet EXP 1 of an expansion valve EXPV and an inlet S 1 of a suction gas heat exchanger SGHX. The inlet SI is in fluid communication with an outlet S4 of the suction gas heat exchanger SGHX, wherein the outlet S4 is in fluid communication with an inlet EXP2 of the expansion valve EXPV.

The expansion valve EXPV is in fluid communication with an inlet El of an evaporator EVAP, and the inlets EXP 1 and EXP2 can be individually controlled to decide whether the refrigerant shall travel directly from the outlet C3 of the condenser COND or via the suction gas heat exchanger SGHX, or both. The inlet El is in fluid communication with an outlet E3 of the evaporator EVAP, and during the passage between the inlet El and the outlet E2, the refrigerant will absorb heat from a heat carrier flowing from an inlet E4 to an outlet E2 of the evaporator EVAP. During its passage between the inlet El and the outlet E3, the refrigerant will evaporate by absorbing heat from the heat carrier flowing between the inlet E4 and the outlet E2.

The outlet E3 is in fluid communication with an inlet S2 of the suction gas heat exchanger SGHX, which in turn is in fluid communication with an outlet S3 of the suction gas heat exchanger SGHX. The outlet S3 is in fluid communication with the compressor C, meaning that the refrigerant circle now is complete.

As is well known by persons skilled in the art, a refrigeration cycle transports heat between a low temperature source and a high temperature source. This is achieved by two different temperature and pressure levels in the refrigeration circuit, namely a high pressure, high temperature zone on a condenser side of the refrigeration circuit between the compressor C and the expansion vale EXPV, and a low temperature, low pressure side on an evaporator side between the expansion valve EXPV and the compressor C.

The purpose of the suction gas heat exchanger SGHX is to allow a heat exchange between cool gaseous refrigerant from the evaporator EVAP and warm liquid refrigerant from the condenser C. This has two effects: 1. The temperature of the gaseous refrigerant exiting the evaporator will be higher, meaning that its density will be lower.

2. The temperature of the liquid refrigerant about to enter the expansion valve will be lower, which will minimize flash boiling of the refrigerant during its passage through the expansion valve.

It might seem that the effects of a suction gas heat exchanger are entirely positive, but this is not the case. If the temperature of the gaseous refrigerant becomes higher, the compressor will have more work compressing the refrigerant, hence reducing the efficiency of the refrigerant cycle.

According to this embodiment, it is possible to control the amount of heat exchange in the suction gas heat exchanger by controlling the inlets EXP1 and EXP2. By closing the inlet EXP2 and opening the inlet EXP1, all warm liquid refrigerant coming from the condenser C will pass the suction gas heat exchanger SGHX and exchange heat with the cold, gaseous refrigerant exiting the evaporator EVAP.

Oppositely, if the inlet EXP2 is open and the inlet EXP1 is closed, all warm liquid refrigerant from the condenser C will go directly from the condenser to the expansion valve EXPV. Hence, there will be no heat exchange between refrigerants in the suction gas heat exchanger SGHX. It is, of course, also possible to control the inlets EXP 1 and EXP2 such that some of the warm liquid refrigerant will pass the suction gas heat exchanger and some will go directly from the condenser C to the expansion valve EXPV.

It should be noted that the two inlets EXP 1 and EXP2 may be separate expansion valves.

A second embodiment is shown in Fig. 2. In this embodiment, the expansion valve EXPV is a single expansion valve, i.e. an expansion valve having a single inlet for all liquid refrigerant. Instead of controlling the amount of heat exchange by controlling separate expansion valves, as in the first embodiment, the amount of heat exchange in the suction gas is controlled by a bypass valve BPl, which when opened allows warm liquid refrigerant from the condenser COND to bypass the suction gas heat exchanger, hence reducing the amount of heat exchange in the suction gas heat exchanger.

Oppositely, when the bypass valve BP l is closed, all of the warm liquid refrigerant will pass the suction gas heat exchanger, hence increasing the heat exchange in the suction gas heat exchanger.

It could be understood, that according to the second embodiment, some warm liquid refrigerant will always pass the suction gas heat exchanger, even if the bypass valve BPl is fully open.

Therefore, according to a third embodiment, also shown in Fig. 2, a second bypass valve BP2 is arranged to control the communication between the condenser COND and the suction gas heat exchanger. By closing the bypass valve BP2 when the bypass valve BPl is open, it is secured that all of the liquid warm refrigerant will bypass the suction gas heat exchanger.

A fourth embodiment is shown in Fig. 3. According to this embodiment, all warm liquid refrigerant from the condenser will pass the suction gas heat exchanger SGHX. The amount of heat exchange in the suction gas heat exchanger SGHX is controlled by an expansion valve EXPV3.

As mentioned above, and as known by persons skilled in the art, there are two pressure levels of the refrigerant in a normal refrigeration cycle - a high pressure level and a low pressure level. According to the third embodiment, however, there will be three different pressure levels, each representing a certain refrigerant boiling temperature:

1. The high pressure level, reaching from the compressor C to the EXPV3. In the high pressure level, the boiling point of the refrigerator will be such that it may condense under heat exchange with the high temperature heat source in the condenser COND.

2. The mid pressure level, reaching from the expansion valve EXPV3 to the

expansion valve EXPV (i.e. the "normal" expansion valve), wherein the boiling point of the refrigerant is controlled such that a desired heat exchange between liquid refrigerant from the condenser C and the gaseous refrigerant from the evaporator EVAP is achieved in the suction gas heat exchanger SGHX.

3. The low pressure level, reaching from the expansion valve EXPV to the

compressor C, wherein the refrigerant pressure level is such that the refrigerant has a boiling point slightly lower than the temperature of the low temperature heat source with which the refrigerant exchanges heat in order to vaporize.

It should be noted that there will be some evaporation of the refrigerant in the mid pressure level - otherwise, its temperature will not decrease. However, due to the cooling in the suction gas heat exchanger SGHX, the evaporated refrigerant will condense again, such that fully liquid refrigerant will enter the expansion valve EXPV.

According to the present invention, it is possible to control the amount of heat exchange between cold gaseous refrigerant about to enter the compressor and warm liquid refrigerant about to enter the expansion valves. A refrigeration cycle is an extremely complicated process - basically, whatever modification you make, it will affect the entire process. However, there are some basic factors:

1. A high ratio between the high pressure side and the low pressure side will allow for a large temperature difference between the source where heat is collected and the temperature source to which the heat is delivered.

2. A high ratio between the high pressure side and the low pressure side will give a low efficiency.

3. Virtually all types of compressors are extremely sensitive to liquid refrigerant. It is crucial that all refrigerant entering the compressor is gaseous.

4. The expansion valves (which, together with the compressor, control the pressure ratio between the high pressure side and the low pressure side) will be impossible to control if the refrigerant passing there through contains any gas at all.

Hence, an optimum running condition for a refrigeration cycle collecting heat from a first heat source having a specific temperature and delivering such heat to a second heat source having higher specific temperature will be roughly as follows:

1. The pressure ratio between the high pressure side and the low pressure side should be such that the refrigerant boiling point on the low pressure side is identical to the temperature of the low temperature heat source. 2. The pressure ratio between the high pressure side and the low pressure side should be such that the refrigerant boiling temperature on the high pressure side is identical to the temperature of the high temperature heat source.

For practical reasons, this is not possible, since the heat transfer in the evaporator and the condenser would be very low. Hence, the pressure ratio must be slightly higher than the theoretically optimal temperature, i.e. such that the refrigerant boiling temperature on the low pressure side is lower than the low temperature heat source and the refrigerant boiling temperature on the high pressure side is higher than the temperature of the high temperature heat source.

Except from causing some efficiency decrease, the difference between the low pressure refrigerant boiling temperature and the low temperature heat source may cause overheating of the gaseous refrigerant leaving the evaporator. As mentioned above, it is crucial that no liquid refrigerant enters the compressor, but overheating of the gaseous refrigerant is detrimental to the compressor efficiency.

By the controllable suction gas heat exchanger presented herein, it is possible to control the overheating of the gaseous refrigerant, hence making it possible to fine- tune the pressure ratio to an optimum level, while securing that the overheating is neither too extensive - causing bad compressor efficiency - nor too low - causing refrigerant droplets entering the compressor, hence risking total compressor failure.

It should be noted that it is required that the expansion valves and the bypass valves used in the different embodiments of the invention are controllable, either in an analogue manner, i.e. by provision of pressure, temperature or pressure ratio sensing means being connected to the valve, e.g. by a capillary tube connected to a bulb in thermal contact with the refrigerant, the temperature of which should control the bypass or expansion valve.

Alternatively, by electronic or digital control of the bypass valves and expansion valves, it is possible to control the amount of heat exchange in the suction gas heat exchanger based on virtually any measured value, e.g. pressure ratio, temperature, compressor speed, load, etc, etc.

By controlling the expansion valves and/or bypass valves, it is possible to optimize a refrigeration system presented herein such that it performs as well at part load as it does on full load. In general terms, the bypass and expansion valves will be controlled such that little or no heat exchange takes place in the suction gas heat exchanger during part load conditions, whereas they will be controlled such that the heat exchange in the suction gas heat exchanger will be significant during full load conditions.

It should be noted that even if the refrigeration system could be used for any kind of refrigeration systems, it is primarily interesting for systems comprising a frequency controlled compressor.