Refrigeration system and method

Abstract

A refrigeration system includes: a compressor for compressing a gaseous refrigerant, causing its temperature and pressure to increase while its boiling point to decrease; a condenser in which the gaseous refrigerant from the compressor exchanges heat with a high-temperature heat carrier, the heat exchange causing the refrigerant to condense; an expansion valve that reduces the pressure of the liquid refrigerant from the condenser, thereby reducing the boiling point of the refrigerant; an evaporator in which the low-boiling-point refrigerant exchanges heat with a low-temperature heat carrier, causing the refrigerant to evaporate; and a suction gas heat exchanger that exchanges heat between the high-temperature liquid refrigerant from the condenser and the low-temperature gaseous refrigerant from the evaporator. The low-temperature gaseous refrigerant entering the suction gas heat exchanger contains a certain amount of low-temperature liquid refrigerant, which evaporates through heat exchange with the high-temperature liquid refrigerant from the condenser. A refrigeration method is also disclosed.

Description

Technical Field

[0001] The present invention relates to a refrigeration system comprising: a compressor for compressing a gaseous refrigerant so that its temperature and pressure increase, wherein its boiling point increases; a condenser in which the gaseous refrigerant from the compressor exchanges heat with a high-temperature heat carrier, said heat exchange causing the refrigerant to condense; an expansion valve which reduces the pressure of the liquid refrigerant from the condenser, thereby reducing the boiling point of the refrigerant; an evaporator in which a low-boiling-point refrigerant exchanges heat with a low-temperature heat carrier, causing the refrigerant to evaporate; and a suction gas heat exchanger which exchanges heat between the high-temperature liquid refrigerant from the condenser and the low-temperature gaseous refrigerant from the evaporator.

[0002] A heat exchanger and a refrigeration system and method are also disclosed. Background Art

[0003] Various brazed plate heat exchangers having a pressed corrugation pattern with ridges and grooves in a herringbone pattern are known in the prior art.It is also known to provide a heat exchanger with an integrated suction gas heat exchanger and to use such a heat exchanger in a refrigeration system.

[0004] In the field of refrigeration, efforts are constantly underway to develop more efficient systems. In practice, optimal refrigeration systems approach the Carnot efficiency, the theoretical upper limit of a heat engine. Generally speaking, any refrigeration system that converts mechanical energy into a temperature difference consists of a compressor, a condenser, an expansion valve, an evaporator, and piping that enables the refrigerant to travel between the compressor, condenser, expansion valve, and evaporator, with heat being transferred from the evaporator to the condenser.

[0005] However, while the efficiency at some temperature differentials may approach the Carnot efficiency, this is far from being true for all operating conditions.

[0006] Generally speaking, all heat exchangers included in a refrigeration system should be as large and efficient as possible. In addition, they should have the lowest possible hold-up volume and low pressure drop. It is understood that these criteria cannot all be met.

[0007] When the temperature after the evaporator is reached, every temperature increase above the temperature at which all the refrigerant is evaporated (i.e. the maximum boiling point of the refrigerant) will mean a loss of efficiency. However, since liquid refrigerant entering the compressor can seriously damage the compressor, it is also crucial that all the refrigerant is actually evaporated before entering the compressor. Although the temperature of the refrigerant does not exceed the boiling temperature, the state in which all the refrigerant is evaporated is often called "zero superheat" and is a very beneficial state in terms of efficiency.

[0008] One way to achieve "zero superheat" in the evaporator is to "flood" the evaporator with liquid refrigerant and allow the refrigerant to evaporate from the flooded evaporator. This configuration is common in large chiller applications, i.e., heat engines with a power of 500-1000 kW. Typically, so-called "plate and shell" or "shell and tube" heat exchangers are used for such applications.

[0009] As can be seen from the above, these evaporator designs offer high performance, but they are far from without drawbacks. First, all heat exchangers, including the housing, are bulky, which means that the material costs for their manufacture are high. Second, and even more importantly, the volume of refrigerant required to fill the heat exchanger is relatively large. In addition to this cost issue, legislation often prohibits excessive refrigerant quantities in heat engines.

[0010] By far the most efficient type of heat exchanger in terms of heat transfer / material quality is the compact brazed plate heat exchanger (BPHE). As known to those skilled in the art, this type of heat exchanger comprises a plurality of plates made from sheet metal and provided with a pressed pattern of ridges and grooves adapted to hold the plates at a distance from one another while forming inter-plate flow channels for the medium to exchange heat. The plates are brazed to one another, which means that each plate pair will be efficient in terms of containing the refrigerant under pressure in the heat exchanger. The advantage of a brazed plate heat exchanger is that virtually all of the material in the heat exchanger contributes to the heat exchange, unlike heat exchangers comprising a shell, the sole purpose of which is to contain the refrigerant.

[0011] The evaporation process in a BPHE differs significantly from that in a flooded shell-and-tube heat exchanger. As mentioned above, evaporation in a flooded shell-and-tube heat exchanger is similar to boiling in a pool. In a BPHE, however, the refrigerant travels more or less linearly through the interplate flow channels. The amount of liquid refrigerant decreases as it approaches the outlet. Due to the increased volume caused by evaporation, the velocity, and therefore the flow resistance, increases along the length of the heat exchanger.

[0012] As mentioned above, it is crucial that no liquid refrigerant enters the compressor. Therefore, it is common for at least some heat exchangers to contain only gaseous refrigerant. The gaseous refrigerant will absorb heat and become unnecessarily hot, which will reduce system efficiency.

[0013] It is also beneficial if the liquid refrigerant entering the evaporator is cold, because flash boiling can be minimized if the refrigerant is cold.

[0014] One way of ensuring a low refrigerant temperature for the refrigerant that is about to enter the expansion valve (thereby reducing the risk of flash boiling) while ensuring a sufficiently high temperature for the gaseous refrigerant that is about to enter the compressor is to use a so-called suction gas heat exchanger. In its simplest form, a suction gas heat exchanger can be made by simply placing the pipe from the evaporator to the compressor close to the pipe from the condenser to the expansion valve and brazing or soldering them together so that heat can be transferred between the pipes. However, for larger systems, it is more common to provide a heat exchanger that is more efficient than simply placing two pipes next to each other. Typically, when using larger types of suction gas heat exchangers, problems with the evaporator outlet pressure drop and the suction gas heat exchanger inlet / outlet pressure drop are destructive to the overall efficiency and may lead to control problems for the system having such a heat exchanger.

[0015] If superheating of the refrigerant can be kept to a minimum while ensuring that no liquid refrigerant enters the compressor, the BPHE can compete with the flooded shell and tube heat exchanger in terms of efficiency while retaining its benefits in terms of compactness and material efficiency.

[0016] In refrigeration technology, so-called "suction heat exchange" is a method for improving the stability of, for example, refrigeration systems. Simply put, suction heat exchange is achieved by exchanging heat between the hot, liquid, high-pressure refrigerant exiting the condenser and the cold, gaseous refrigerant exiting the evaporator. Through suction heat exchange, the temperature of the cold, gaseous refrigerant increases, while the temperature of the hot liquid decreases. This has two positive effects: first, the problem of flash evaporation after the hot liquid has passed through the subsequent expansion valve is reduced; second, the risk of liquid droplets in the gaseous refrigerant leaving the evaporator is reduced.

[0017] Suction heat exchange is well known. Typically, suction heat exchange can be achieved by simply brazing or soldering the pipes carrying the refrigerant in the desired heat exchange state. However, this method of achieving heat exchange is expensive in terms of the refrigerant volume required—it is always beneficial if the pipes between different components of the refrigeration system are as short as possible. Suction heat exchange by brazing or soldering pipes carrying fluids at different temperatures requires longer pipes than would otherwise be the case. Consequently, the internal volume of the pipes increases, requiring more refrigerant in the refrigeration system. This is not only disadvantageous from an economic standpoint, but also because the amount of refrigerant is limited in some jurisdictions.

[0018] Another option is to provide a separate heat exchanger for suction gas heat exchange. A separate heat exchanger is more efficient than simply brazing different piping sections to each other, but providing a separate heat exchanger also requires piping connecting the evaporator and condenser to the suction gas heat exchanger, which increases the refrigerant volume of the refrigeration system.

[0019] Furthermore, refrigeration systems typically need to be able to operate in both heating and cooling modes, depending on the required / desired load. Typically, switching between heating and cooling modes is accomplished by switching a four-way valve so that the evaporator becomes the condenser and the condenser becomes the evaporator. Unfortunately, this means that the heat exchange in either or both of the condenser / evaporator units will be a co-current heat exchange, meaning that the medium exchanging heat travels in the same general direction in either heating or cooling mode. As is well known to those skilled in the art, co-current heat exchange is inferior to counter-current heat exchange. In the evaporator, a reduction in heat exchange performance can lead to an increased risk of droplets in the refrigerant vapor leaving the heat exchanger. Such droplets can severely damage the compressor and are therefore highly undesirable. However, devices for changing the flow direction of the medium to exchange heat with the refrigerant in the evaporator are expensive and increase the complexity of the refrigeration system.

[0020] It is an object of the present invention to solve or at least alleviate the above-mentioned and other problems.

[0021] It is an object of the present invention to provide a plate heat exchanger that provides advantageous fluid distribution and heat transfer between fluids in a refrigeration system.

[0022] Another object of the present invention is to provide a high-efficiency refrigeration system.

[0023] It is another object of the present invention to provide a BPHE and a refrigeration system wherein the BPHE is configured to achieve zero or near zero superheat of the refrigerant entering the compressor. Summary of the Invention

[0024] According to a first aspect of the present invention, some of the above-mentioned purposes are achieved by a refrigeration system, which includes: a compressor for compressing gaseous refrigerant so that the temperature and pressure of the gaseous refrigerant increase, wherein the boiling point of the gaseous refrigerant increases; a condenser, in which the gaseous refrigerant from the compressor exchanges heat with a high-temperature heat carrier, and the heat exchange causes the refrigerant to condense; an expansion valve, which reduces the pressure of the liquid refrigerant from the condenser, thereby reducing the boiling point of the refrigerant; an evaporator, in which the low-boiling-point refrigerant exchanges heat with a low-temperature heat carrier, so that the refrigerant evaporates; and a suction gas heat exchanger, which exchanges heat between the high-temperature liquid refrigerant from the condenser and the low-temperature gaseous refrigerant from the evaporator, characterized in that the balancing valve is arranged to be able to bypass the high-temperature liquid refrigerant so that the high-temperature liquid refrigerant does not exchange heat with the low-temperature gaseous refrigerant from the evaporator in the suction gas heat exchanger.

[0025] The invention also relates to a method for controlling such a system, comprising the following steps

[0026] a) measuring the temperature of the high-temperature liquid refrigerant,

[0027] b) measuring the temperature of the low-temperature gaseous refrigerant,

[0028] c) calculating the temperature difference between the high temperature liquid refrigerant and the low temperature gaseous refrigerant, and,

[0029] d) If the temperature difference is less than a predetermined threshold, controlling the balancing valve to bypass the suction gas heat exchanger.

[0030] For example, the threshold may be zero.

[0031] According to a second aspect of the present invention, some of the above-mentioned purposes are achieved by a refrigeration system, which includes: a compressor for compressing gaseous refrigerant so that the temperature and pressure of the gaseous refrigerant increase, wherein the boiling point of the gaseous refrigerant increases; a condenser, in which the gaseous refrigerant from the compressor exchanges heat with a high-temperature heat carrier, and the heat exchange causes the refrigerant to condense; an expansion valve, which reduces the pressure of the liquid refrigerant from the condenser, thereby reducing the boiling point of the refrigerant; an evaporator, in which the low-boiling-point refrigerant exchanges heat with a low-temperature heat carrier, so that the refrigerant evaporates; and a suction gas heat exchanger, which performs heat exchange between the high-temperature liquid refrigerant from the condenser and the low-temperature gaseous refrigerant from the evaporator, characterized in that the low-temperature gaseous refrigerant entering the suction gas heat exchanger contains a certain amount of low-temperature liquid refrigerant, and the low-temperature liquid refrigerant evaporates due to the heat exchange with the high-temperature liquid refrigerant from the condenser.

[0032] According to a third aspect of the present invention, some of the above-mentioned purposes are achieved by a plate heat exchanger, which includes a plurality of heat exchanger plates, which are provided with a pressing pattern, which is suitable for providing contact points that keep the heat exchanger plates at a certain distance from each other, so that inter-plate flow channels are formed between the plates, and the heat exchanger is provided with inter-plate flow channels, which are used for heat exchange between a first medium and a second medium in the inter-plate flow channels and a third medium in the inter-plate flow channels, wherein the inter-plate flow channels are selectively fluidly connected to port openings for the first medium, the second medium and the third medium, and is characterized in that the first and second integrated suction gas heat exchanger parts are arranged near the port openings for the second medium and the third medium.

[0033] According to a fourth aspect of the present invention, some of the above objects are achieved by a brazed plate heat exchanger comprising a plurality of first and second heat exchanger plates, wherein the first heat exchanger plates are formed with a first pattern of ridges and grooves, and the second heat exchanger plates are formed with a second pattern of ridges and grooves, contact points being provided between at least some of the intersecting ridges and grooves of adjacent plates in forming inter-plate flow channels for fluidly exchanging heat, the inter-plate flow channels being in selective fluid communication with first, second, third, and fourth large port openings and first and second small port openings, wherein the first and second heat exchanger plates form There is a dividing surface dividing the heat exchanger plates into a first heat exchange portion and a second heat exchange portion, so that a fluid passing between the first and second large port openings exchanges heat with a fluid passing between a third and fourth port openings on the first heat exchange portion of each plate and a fluid passing between the first and second small port openings on the second heat exchange portion of each plate, characterized in that the ridges and grooves are formed so that the inter-plate flow channels between different pairs of plates have different volumes, and optionally, the first pattern at least partially presents a first angle, such as a first herringbone angle, and the second pattern at least partially presents a second angle, such as a second herringbone angle different from the first angle.

[0034] Small port openings and dividing surfaces result in an integrated suction gas heat exchanger and, together with different inter-plate flow channel volumes, a BPHE with advantageous properties, for example, for use in refrigeration systems. By combining different herringbone angles and inter-plate flow channel volumes, fluid flow distribution and pressure drop can be balanced to achieve efficient heat exchange, which has been found to be particularly advantageous for refrigeration. This BPHE has been found to result in virtually zero or near-zero superheat for the refrigerant entering the compressor in a refrigeration system. Evaporation results in nearly zero superheat, while superheat increases outside the evaporator relative to the water side (secondary side), increasing superheat and carryover. Carryover droplets evaporate during the suction gas heat exchange process. By reducing heat transfer from the gas to the water / brine in the heat exchanger, superheat does not affect the evaporation process, as would occur with increasing superheat in a standard heat exchanger. This opens the possibility of using co-current and near-temperature approaches.

[0035] The present invention also relates to a refrigeration system and a refrigeration method comprising the plate heat exchanger.

[0036] According to a fifth aspect of the present invention, some of the above objects are achieved by a brazed plate heat exchanger comprising a plurality of first and second heat exchanger plates, wherein the first heat exchanger plates are formed with a first pattern of ridges and grooves, and the second heat exchanger plates are formed with a second pattern of ridges and grooves, contact points being provided between at least some of the intersecting ridges and grooves of adjacent plates in forming inter-plate flow channels for fluidly exchanging heat, the inter-plate flow channels being selectively fluidly connected via port openings, wherein the first pattern of ridges and grooves differs from the second pattern of ridges and grooves such that the volume of the inter-plate flow channels on one side of the first heat exchanger plate is different from the volume of the inter-plate flow channels on the opposite side of the first heat exchanger plate, the first pattern of ridges and grooves being at a first angle, and the second pattern of ridges and grooves being at a second angle different from the first angle.

[0037] The combination of different interplate flow channel volumes on opposite sides of a plate and at least two different plate patterns with different angles results in a BPHE with properties that favor fluid distribution, where fluid flow distribution and pressure drop can be balanced for efficient heat exchange. This enables different characteristics to be achieved in the interplate flow channels on opposite sides of the same plate, where the flow and pressure drop on one side can differ from the opposite side. Furthermore, different flow channel volumes on opposite sides of a plate can be used for different types of media, such as liquid in one channel and gas in the other. Furthermore, the combination of different interplate flow channel volumes in adjacent interplate flow channels and at least two different plate patterns with different angles results in different brazed joint shapes, such as the width of the brazed joint relative to the direction of media flow, to control the flow and pressure drop of the media.

[0038] When a refrigerant begins to evaporate, it transitions from a liquid to a gas. Liquids have a much higher density than vapors. For example, the liquid density of R410A at Tdew = 5°C is 32 times higher than the vapor density. This also means that the vapor will move through the channels at a speed 32 times higher than the liquid. This automatically results in a 32-fold higher dynamic pressure drop for the vapor than for the liquid, meaning that vapor produces a much higher pressure drop for all types of refrigerants.

[0039] The performance of a heat exchanger (Temperature Approach, TA) is defined as the water outlet temperature (at the inlet of the heat exchanger channel) minus the evaporation temperature (Tdew) at the outlet of the heat exchanger channel. High pressure drops along the heat exchanger surface result in different local saturation temperatures, which in turn lead to a relatively large overall difference in refrigerant temperature between the inlet and outlet of the channel. The temperature at the channel inlet will be higher. This has a direct, detrimental impact on the performance of the heat exchanger, as the higher inlet refrigerant temperature (due to excessive channel pressure drop) makes it more difficult to cool the outlet water to the correct temperature. The only way for the system to compensate for the excessively high refrigerant inlet temperature is by lowering the evaporation temperature until the correct water outlet temperature can be achieved. By creating a pattern for heat exchanger channels that have high heat transfer characteristics while also having low pressure drop characteristics, the heat exchanger can achieve higher performance. A lower overall refrigerant pressure drop in the channels not only improves heat exchanger performance, but also has a positive impact on overall system performance and, therefore, energy consumption.

[0040] Also disclosed is the use of brazed plate heat exchangers with different inter-plate flow channel volumes and different angles, with or without a suction gas heat exchanger, for evaporation or condensation of a medium.

[0041] According to a sixth aspect of the present invention, some of the above objects are achieved by a brazed plate heat exchanger comprising a plurality of first and second heat exchanger plates, wherein the first heat exchanger plates are formed with a first pattern of ridges and grooves and the second heat exchanger plates are formed with a second pattern of ridges and grooves, contact points being provided between at least some of the intersecting ridges and grooves of adjacent plates in forming interplate flow channels for fluidly exchanging heat, the interplate flow channels being selectively fluidically connected to port openings, characterized in that the first pattern of ridges and grooves is different from the second pattern of ridges and grooves such that the volume of the interplate flow channels on one side of the first heat exchanger plate is different from the volume of the interplate flow channels on the opposite side of the first heat exchanger plate. Optionally, the first pattern presents a first angle and the second pattern presents a second angle different from the first angle. The heat exchanger is provided with a modified port heat exchanger.

[0042] The present invention also relates to a refrigeration system and a refrigeration method having a heat exchanger having two or more different plates having different patterns and provided with a retrofit port heat exchanger. BRIEF DESCRIPTION OF THE DRAWINGS

[0043] The present invention will now be described with reference to the accompanying drawings, in which:

[0044] Figure 1 is an exploded perspective view of a heat exchanger according to one embodiment of the present invention;

[0045] Figure 2 yes Figure 1 an exploded perspective view of a portion of a heat exchanger illustrating a first heat exchanger plate and a second heat exchanger plate of the heat exchanger;

[0046] Figure 3 is a schematic cross-sectional view of another portion of a first heat exchanger plate according to one embodiment, showing grooves of the same depth of the first heat exchanger plate;

[0047] Figure 4 is a schematic cross-sectional view of a portion of a second heat exchanger plate according to one embodiment, illustrating optional depths of the grooves of the second heat exchanger plate;

[0048] Figure 5 is a schematic cross-sectional view of a portion of a heat exchanger including first and second heat exchanger plates according to one embodiment, wherein the first and second heat exchanger plates are arranged alternately;

[0049] Figure 6a is a schematic front view of a first heat exchanger plate showing a corrugated herringbone pattern thereof having a first herringbone angle according to one embodiment;

[0050] Figure 6b is a schematic elevational view of a first heat exchanger plate according to an alternative embodiment, showing a corrugation pattern thereof having a first angle;

[0051] Figure 7a is a schematic front view of a second heat exchanger plate according to one embodiment, showing a corrugated herringbone pattern having a second herringbone angle;

[0052] Figure 7b is a schematic elevational view of a second heat exchanger plate according to an alternative embodiment, showing a corrugation pattern thereof having a second angle;

[0053] Figure 8 is a schematic diagram of a first heat exchanger plate disposed on a second heat exchanger plate, showing Figure 5 instances of the points of contact between them;

[0054] Figure 9 is a schematic diagram of a second heat exchanger plate disposed on a first heat exchanger plate, showing Figure 5 instances of the points of contact between them;

[0055] Figure 10a and Figure 10b is a schematic plan view showing a refrigeration system in a heating mode according to various embodiments of the present invention;

[0056] Figure 11a and Figure 11bIt shows that according to Figure 10a and Figure 11a A schematic plan view of a refrigeration system in cooling mode;

[0057] Figure 12 is an exploded perspective view of a heat exchanger to be mated with a retrofit port heat exchanger according to one embodiment of the present invention;

[0058] Figure 13 is a schematic perspective view of a retrofit port heat exchanger according to one embodiment;

[0059] Figure 14 is a schematic perspective view of a retrofit port heat exchanger according to an alternative embodiment;

[0060] Figure 15 is a schematic cross-sectional view of a portion of a heat exchanger including first and second heat exchanger plates according to another embodiment;

[0061] Figure 16 is a schematic cross-sectional view of a portion of a heat exchanger including first and second heat exchanger plates according to another embodiment;

[0062] Figure 17 is a schematic cross-sectional view of a portion of a heat exchanger including first and second heat exchanger plates according to yet another embodiment;

[0063] Figure 18 is a schematic cross-sectional view of a portion of a stack of first and second heat exchanger plates having different corrugation depths according to another embodiment;

[0064] Figure 19 is a schematic exploded perspective view of a true dual heat exchanger including a dual integrated suction gas heat exchanger according to one embodiment of the present invention; and

[0065] Figure 20 is a schematic perspective view of another embodiment of a corrugation pattern of a heat exchanger plate, showing a corrugation pattern wherein the angle of the corrugation pattern in the central main heat exchanging portion is different from the angle in the portion at the port openings of the heat exchanger plate. DETAILED DESCRIPTION

[0066] Reference Figure 1 , according to one embodiment, a brazed plate heat exchanger 100 is shown, wherein a portion thereof is Figure 2, which is shown in more detail in FIG. The heat exchanger 100 includes a plurality of first heat exchanger plates 110 and a plurality of second heat exchanger plates 120 stacked together to form the heat exchanger 100. The first and second heat exchanger plates 110, 120 are arranged alternately, with every other plate being a first heat exchanger plate 110 and every other plate being a second heat exchanger plate 120. Alternatively, the first and second heat exchanger plates can be arranged in another configuration along with additional heat exchanger plates. The heat exchanger 100 is an asymmetric plate heat exchanger.

[0067] The heat exchanger plates 110, 120 are made of sheet metal and are provided with a pressed pattern of ridges R1, R2a, R2b and grooves G1, G2a, G2b so that when the plates are stacked together to form the heat exchanger 100, inter-plate flow channels for fluids to exchange heat are formed between the plates by providing contact points between at least some of the intersecting ridges and grooves of adjacent plates 110, 120, whereby inter-plate flow channels for fluids to exchange heat are formed. Figure 1 and Figure 2 The pressed pattern is a herringbone pattern. However, the pressed pattern can also be in the form of straight lines extending at an angle. In any case, the pressed pattern of ridges and grooves is a wavy pattern. Except for the contact points, the pressed pattern is suitable for maintaining a certain distance between the plates 110 and 120.

[0068] In the embodiment shown, each of the heat exchanger plates 110 , 120 is surrounded by a skirt S extending generally perpendicular to the plane of the heat exchanger plate and adapted to contact the skirt of an adjacent plate to provide a seal along the perimeter of the heat exchanger 100 .

[0069] Heat exchanger plates 110 and 120 are arranged with large port openings O1-O4 and small port openings SO1 and SO2 for transferring heat between fluids and the inter-plate flow channels. In the illustrated embodiment, heat exchanger plates 110 and 120 are arranged with a first large port opening O1, a second large port opening O2, a third large port opening O3, and a fourth large port opening O4. Furthermore, heat exchanger plates 110 and 120 are arranged with a first small port opening SO1 and a second small port opening SO2. The areas surrounding the large port openings O1 to O4 are positioned at different heights, thereby enabling selective communication between the large port openings and the inter-plate flow channels. In heat exchanger 100, the areas surrounding the large port openings O1-O4 are arranged such that the first and second large port openings O1 and O2 are in fluid communication with each other via some inter-plate flow channels, while the third and fourth large port openings O3 and O4 are in fluid communication with each other via adjacent inter-plate flow channels. In the illustrated embodiment, the heat exchanger plates 110, 120 are rectangular with rounded corners, with the large port openings O1-O4 arranged near the corners. Alternatively, the heat exchanger plates 110, 120 are square, which may, for example, have rounded corners. Alternatively, the heat exchanger plates 110, 120 are circular, oval, or arranged in another suitable shape, with the large port openings O1-O4 distributed in a suitable manner. In the illustrated embodiment, each heat exchanger plate 110, 120 is formed with four large port openings O1-O4. In other embodiments of the present invention, as described below, the number of large port openings may be greater than four, namely, six, eight, or ten. For example, the number of large port openings is at least six, wherein the heat exchanger is configured to provide heat exchange between at least three fluids. Thus, according to one embodiment, the heat exchanger is a three-circuit heat exchanger having at least six large port openings, and may or may not be provided with at least one integrated suction gas heat exchanger.

[0070] In the illustrated embodiment, each of the heat exchanger plates 110, 120 is formed with two small port openings SO1, SO2. The small port openings SO1, SO2 are arranged to provide an integrated suction gas heat exchanger. Thus, the first and second heat exchanger plates 110, 120 are formed with a dividing surface DW that divides the heat exchanger plates 110, 120 into a first heat exchange portion 130 and a second heat exchange portion 140, so that fluid passing between the first and second large port openings O1, O2 exchanges heat with fluid passing between the third and fourth port openings O3, O4 in the first heat exchange portion 130 of each plate 110, 120, and with fluid passing between the first and second small port openings SO1, SO2 in the second heat exchange portion 140 of each plate 110, 120.

[0071] A dividing surface DW is provided to divide the heat exchange area into the first heat exchange section 130 and the second heat exchange section 140. For example, the dividing surface DW is arranged between a long side and an adjacent short side of the heat exchanger plates 110, 120. For example, the dividing surface DW extends from the long side to the short side. Alternatively, the dividing surface DW is arranged between the two long sides, for example, extending from one long side to the other long side. In the illustrated embodiment, the dividing surface DW curves between the long and short sides of the plates. Alternatively, the dividing surface DW is straight or formed with an angle.

[0072] Separation surface DW comprises elongated flat surfaces disposed at different heights on different plates 110 and 120. When the flat surfaces of adjacent plates 110 and 120 contact each other to form separation surface DW, the inter-plate flow channel is sealed; if they are not in contact, the channel remains open. In this case, separation surface DW is disposed at the same height as the area surrounding the first and second large port openings O1 and O2. This means that separation surface DW is open to the inter-plate flow channel fluidically connecting the first and second large port openings O1 and O2, while blocking the fluid in the inter-plate flow channel fluidically connecting the third and fourth large port openings O3 and O4.

[0073] Because the dividing surface DW blocks fluid flow in the interplate flow channels connected to the third and fourth large port openings O3 and O4, separate interplate flow channels exist on either side of the dividing surface DW. On one side of this dividing surface DW, the interplate flow channels not connected to the third and fourth large port openings O3 and O4 communicate with the two small port openings SO1 and SO2. It should be noted that the dividing surface DW does not block the interplate flow channels connected to the first and second large port openings O1 and O2; therefore, the medium flowing in the interplate flow channels connected to the small port openings SO1 and SO2 exchanges heat with the medium flowing in the flow channels connected to the first and second large port openings O1 and O2—just as the medium flowing in the interplate flow channels connected to the third and fourth large port openings O3 and O4 does.

[0074] exist Figure 1 and Figure 2In the illustrated embodiment, the dividing surface DW extends between the first large port opening O1 and the third large port opening O3. Small openings SO1 and SO2 are located on either side of the first large port opening O1. It should be noted that the first large port opening O1 is arranged so that the medium flowing in the inter-plate flow channel communicating with the small port openings SO1 and SO2 can pass through both sides of the first large port opening O1. The dividing surface DW extends between the first large port opening O1 and the remaining large port openings O2-O4. The first and second small openings SO1 and SO2 are located on the same side of the dividing surface DW as the first large port opening O1, i.e., in the second heat exchange section 140. The remaining large port openings O2-O4 are arranged on the other side of the dividing surface DW, i.e., outside the dividing wall DW and in the first heat exchange section 130.

[0075] In the embodiment shown, the heat exchanger 100 comprises only a first heat exchanger plate 110 and a second heat exchanger plate 120. Alternatively, the heat exchanger 100 comprises a third heat exchanger plate and optionally also a fourth heat exchanger plate, wherein the third heat exchanger plate and the optional fourth heat exchanger plate are arranged with a different pressing pattern than the first heat exchanger plate 110 and the second heat exchanger plate 120, and wherein these heat exchanger plates are arranged in a suitable sequence.

[0076] In the illustrated embodiment, the heat exchanger 100 further includes a start plate 150 and an end plate 160. The start plate 150 is formed with openings corresponding to the large port openings O1-O4 and the small port openings SO1 and SO2 for allowing fluid to flow into and out of the inter-plate flow channels formed by the first heat exchanger plates 110 and the second heat exchanger plates 120. For example, the end plate 160 is a conventional end plate.

[0077] Reference Figure 3 , schematically shows a cross-sectional view of a first heat exchanger plate 110 according to one embodiment. The first heat exchanger plate 110 is formed with a first pattern of ridges R1 and grooves G1. The grooves G1 of the first heat exchanger plate are formed to have the same depth D1. Figure 3. Thus, all grooves G1 are formed to have the same depth D1. For example, the depth D1 is 0.5-5 mm, such as 0.6-3 mm or 0.8-3 mm. For example, all ridges R1 are formed in a corresponding manner to have the same height. In other words, the corrugation depth of the first heat exchanger plate 110 is symmetrical and similar across the entire plate, or at least substantially across the entire plate. According to one embodiment, at least the first heat exchange portion 130 of the first heat exchanger plate 110 (e.g., the entire first heat exchange portion 130 thereof) is formed with the same corrugation depth, wherein each groove G1 is formed to have a depth D1. For example, the first heat exchange portion 130 and the second heat exchange portion 140 of the first heat exchanger plate 110 (e.g., the entire first heat exchange portion 130 and the entire second heat exchange portion) are formed with the same corrugation depth, wherein each groove G1 is formed to have a depth D1.

[0078] Reference Figure 4 , schematically illustrates a cross-sectional view of a second heat exchanger plate 120 according to one embodiment. For example, all second heat exchanger plates 120 are identical. The second heat exchanger plates 120 are formed with a second pattern of first and second ridges R2a, R2b and first and second grooves G2a, G2b. The first and second grooves G2a, G2b of the second heat exchanger plate 120 are formed to have different depths, with the first groove G2a having a first depth D2a and the second groove G2b having a second depth D2b, wherein the second depth D2b is different from the first depth D2a. For example, the first depth D2a is 0.5-5 mm, such as 0.6-3 mm or 0.8-3 mm, while the second depth D2b is 30-80% of the first depth D2a, such as 40-60% thereof. Similarly, the ridges R2a, R2b have different heights. In the illustrated embodiment, the first depth D2a is greater than the second depth D2b. The first and second grooves G2a, G2b are arranged alternately. Alternatively, the first and second grooves G2a, G2b and optionally other grooves having other depths are arranged in any desired pattern.

[0079] For example, the pattern of ridges and grooves of the second heat exchanger plate 120 is asymmetric, ie the second heat exchanger plate 120 forms an asymmetric heat exchanger when combined with the first heat exchanger plate 110, for example as described below with reference to Figure 5According to one embodiment, at least the first heat exchange portion 130 of the second heat exchanger plate 120 (e.g., the entire first heat exchange portion 130 thereof) is formed with a second pattern of ridges and grooves, wherein the second pattern of ridges and grooves has grooves of at least two different corrugation depths D2a and D2b. For example, the first heat exchange portion 130 and the second heat exchange portion 140 of the first heat exchanger plate 110 (e.g., the entire first heat exchange portion 130 and the entire second heat exchange portion) are formed with at least two corrugation depths, wherein the first grooves G2a are formed to have a first depth D2a, and the second grooves G2b are formed to have a second depth D2b.

[0080] Reference Figure 5 A plurality of first and second heat exchanger plates 110, 120 have been stacked to schematically illustrate the formation of inter-plate flow channels according to one embodiment. In the illustrated embodiment, every other plate is a first heat exchanger plate 110, and the remaining plates are second heat exchanger plates 120. The first and second heat exchanger plates are arranged alternately to form an asymmetric heat exchanger 100, wherein the inter-plate flow channels are formed with different volumes. Alternatively, the different volumes of the inter-plate flow channels are formed by extending the profile at the same pressing depth or corrugation depth. For example, the first and second heat exchanger plates are provided with different corrugation depths. For example, the first and / or second heat exchanger plates are asymmetric heat exchanger plates. Alternatively, the first and / or second heat exchanger plates are symmetric heat exchanger plates.

[0081] Reference Figure 6a, schematically illustrating a first pattern of ridges R1 and grooves G1 of a first heat exchanger plate 110. The pattern is a pressed herringbone pattern, wherein the ridges R1 and grooves G1 are arranged with two inclined legs that meet at a vertex, for example, at a centrally located vertex, to form an arrowhead shape. For example, the vertices are distributed along an imaginary centerline (e.g., the longitudinal centerline of a rectangular heat exchanger plate). For example, the herringbone pattern is arranged such that the ridges R and grooves G, at least in the central portion of the first heat exchanger plate 110, extend from one long side of the first heat exchanger plate 110 to the other long side, for example, with all vertices pointing toward one of the short sides. The pattern of the first heat exchanger plate 110 (i.e., the first pattern of ridges R1 and grooves G1) exhibits a first herringbone angle β1. This herringbone angle is the angle between the ridges and an imaginary line passing through the plate, perpendicular to the long sides of the rectangular plate, schematically illustrated by the dashed line C. Thus, the herringbone angle is the angle between the ridges and the short sides of the heat exchanger plate, with the vertices pointing toward the short sides. The long sides of the heat exchanger plates extend perpendicular to the short sides, and therefore, the ridge and groove pattern is also arranged so that the ridges are angled relative to the long sides. For example, the herringbone angle is the same on both sides of the vertex. For example, the entire or substantially the entire first pattern of ridges and grooves can be formed with a first herringbone angle β1 across the entire plate, or at least across the first heat exchange portion 130 and, for example, also across the second heat exchange portion 140. For example, the first herringbone angle β1 is 25° to 70° or 30° to 45°.

[0082] Reference Figure 6b According to an alternative embodiment, a first pattern of ridges R1 and grooves G1 of a first heat exchanger plate 110 is schematically shown, wherein the embossed pattern is in the form of obliquely extending straight lines. Thus, the embossed pattern of ridges and grooves is a corrugated pattern of obliquely extending straight lines. The obliquely extending straight lines of the first heat exchanger plate 110 are arranged at an angle β1. For example, the pattern is arranged such that the ridges R1 and grooves G1 extend parallel from one long side of the first heat exchanger plate 110 to the other long side.

[0083] Reference Figure 7a, schematically illustrating a second pattern of ridges R2a, R2b and grooves G2a, G2b of the second heat exchanger plate 120. This second pattern is a pressed herringbone pattern as described above with reference to the first heat exchanger plate 110, but with a second herringbone angle β2 that is different from the first herringbone angle β1. Thus, the second heat exchanger plate 120 is arranged in a herringbone pattern having a different angle than the first heat exchanger plate. For example, the second herringbone angle β2 is 0° to 90°, 25° to 70°, or 30° to 45°. For example, the entire or substantially the entire pattern of ridges and grooves of the second heat exchanger plate 120 is formed with the second herringbone angle β2 throughout the entire plate, or at least throughout the first heat exchange portion 130 and, for example, also throughout the second heat exchange portion 140. For example, the difference between the first herringbone angle β1 and the second herringbone angle β2 is 2° to 35°.

[0084] Reference Figure 7b According to an alternative embodiment, a second pattern of ridges R2a, R2b and grooves G2a, G2b of the second heat exchanger plate 120 is schematically shown, wherein the embossed pattern is in the form of obliquely extending straight lines. Thus, the embossed pattern of ridges and grooves is a corrugated pattern of obliquely extending straight lines. The obliquely extending straight lines of the second heat exchanger plate 120 are arranged at an angle β2. For example, the pattern is arranged such that the ridges R2a, R2b and grooves G2a, G2b extend parallel from one long side of the second heat exchanger plate 120 to the other long side.

[0085] Therefore, the first and second heat exchanger plates 110 and 120 are formed with different herringbone angles β1 and β2 and different pressing patterns, resulting in different inter-plate volumes. For example, the first and second heat exchanger plates 110 and 120 are provided with different corrugation depths. Alternatively or additionally, the first and second heat exchanger plates 110 and 120 are provided with different corrugation frequencies. For example, the first and second heat exchanger plates 110 and 120 have the same corrugation depth but different corrugation frequencies. Thus, the first and second heat exchanger plates 110 and 120 are provided with different corrugation depths and / or different corrugation frequencies. For example, one of the first and second heat exchanger plates 110 and 120 is a symmetrical heat exchanger plate, while the other is asymmetrical. Alternatively, both the first and second heat exchanger plates 110 and 120 are asymmetrical. Alternatively, both the first and second heat exchanger plates 110 and 120 are symmetrical.

[0086] exist Figure 8 and Figure 9 In, use Figure 5 The example of FIG schematically illustrates the contact points between the first plate 110 and the second plate 120. A brazed joint 170 is formed in and / or around the contact points 170 between the intersecting ridges and grooves. Figure 8 and Figure 9 In the embodiment of the present invention, the solder joints 170 are formed in all contact points. Alternatively, the solder joints 170 are formed only in some contact points. Figure 8 In the embodiment, the first heat exchanger plate 110 is arranged on the second heat exchanger plate 120, wherein the contact points are formed in a first pattern. Figure 8 In the embodiment, all intersections between the ridges R1 of the first heat exchanger plate 110 and the ridges or grooves of the second heat exchanger plate 120 serve as contact points.

[0087] Figure 9 is a schematic diagram of a second heat exchanger plate 120 arranged on the first heat exchanger plate 110, wherein the contact points are formed in a second pattern. Figure 9 In the embodiment, only the intersections between the first ridges R2a of the second heat exchanger plates 120 serve as contact points that can form brazing joints 170, wherein the second ridges R2b are arranged with gaps from the intersecting ridges or grooves of the first heat exchanger plates 110. Therefore, no contact points are formed between the second ridges R2b of the second heat exchanger plates 120 and the first heat exchanger plates 110, and no brazing joints are formed. Figure 9 In FIG. 1 , all contact points are shown as solder joints 170 .

[0088] According to one embodiment, the brazed joints 170 between the first and second heat exchanger plates 110, 120 are elongated, for example, elliptical in shape. The brazed joints 170 are arranged along a first direction in the inter-plate flow channels with the larger volume, and along a second direction in the inter-plate flow channels with the smaller volume, to provide a desired pressure drop in the inter-plate flow channels. For example, the brazed joints 170 are arranged at a first angle relative to the longitudinal direction of the plates 110, 120 in the inter-plate flow channels with the larger volume, and at a second angle in the remaining inter-plate flow channels. According to one embodiment, the first angle is greater than the second angle.

[0089] exist Figure 10a 、 Figure 10b and Figure 11a 、 Figure 11b , an embodiment of a chiller system in heating mode and cooling mode is shown respectively, and the chiller system can use the heat exchanger 100 according to any one of the above heat exchanger embodiments. The chiller system can also be called a refrigeration system.

[0090] according to Figure 10a 、 Figure 10b 、 Figure 11a 、 Figure 11bThe chiller system of this embodiment includes a compressor C, a four-way valve FWV, a payload heat exchanger PLHE connected to a brine system requiring heating or cooling, a first controllable expansion valve EXPV1, a first check valve OWV1, a dump heat exchanger DHE connected to a heat source capable of dumping unwanted heat or cold, a second expansion valve EXPV2, and a second check valve OWV2. Each of the heat exchangers PLHE and DHE has four large openings O1-O4 and two small openings SO1 and SO2, as described above. The large openings O1 and O2 of each heat exchanger are interconnected, the large openings O3 and O4 of each heat exchanger are interconnected, and the small openings SO1 and SO2 of each heat exchanger are interconnected. Heat exchange occurs between the fluid flowing from O1 to O2 and the fluids flowing between O3 and O4, and between SO1 and SO2. However, heat exchange does not occur between the fluid flowing from O3 to O4 and the fluid flowing from SO1 to SO2. The payload heat exchanger PLHE and / or the dump heat exchanger DHE is a plate heat exchanger 100 as described herein.

[0091] In heating mode, if Figure 10a and Figure 10b As shown, compressor C delivers high-pressure gaseous refrigerant to four-way valve FWV. In this heating mode, the four-way valve is controlled to direct the high-pressure gaseous refrigerant to the large opening O1 of the payload heat exchanger PLHE. The high-pressure gaseous refrigerant then passes through the payload heat exchanger PLHE and exits at the large opening O2. While passing through the payload heat exchanger PLHE, the high-pressure gaseous refrigerant exchanges heat with the brine solution connected to the payload, requiring heating, and flows from large opening O4 to large opening O3, countercurrently to the refrigerant flowing from large opening O1 to large opening O2. While exchanging heat with the brine solution, the high-pressure gaseous refrigerant condenses, and upon exiting the payload heat exchanger PLHE through the large opening O2, it is fully condensed, i.e., in a liquid state.

[0092] In heating mode, the first expansion valve EXPV1 will be fully closed and the liquid refrigerant flow leaving the payload heat exchanger will pass through the first one-way valve OWV1, which allows refrigerant flow in one direction while preventing flow in the other direction (this will be explained later in conjunction with the description of cooling mode).

[0093] After having passed the first non-return valve OWV1, the liquid refrigerant (still relatively hot) will enter the small opening SO2 of the dump heat exchanger DHE and leave the heat exchanger through the small opening SO1. During the passage between the small openings SO and SO1, the temperature of the refrigerant drops significantly due to the heat exchange with the cold, mainly gaseous refrigerant that will leave the dump heat exchanger DHE.

[0094] During a cold start, for example, before the system reaches favorable operating conditions, it may be necessary to balance the heat exchange capacity in the suction gas heat exchanger. This can be achieved by controlling a balancing valve BV, such as a three-way valve, which is arranged to control the flow of liquid refrigerant from the condenser to either or both of the small opening SO2 and the expansion valve EXPV2, thereby controlling the heat exchange capacity in the suction gas heat exchanger.

[0095] After exiting the dump heat exchanger DHE through the small opening SO1, the liquid refrigerant passes through the second expansion valve EXPV2, where its pressure drops, causing some refrigerant to flash, resulting in a drop in temperature. From the second expansion valve, the refrigerant passes through a branch connected to the second one-way valve OWV2. This branch connects the high-pressure and low-pressure sides of the refrigerant circuit and is closed to refrigerant flow due to the pressure difference between the high-pressure and low-pressure sides. After passing through the branch, the cold, low-pressure, semi-liquid refrigerant enters the large opening O2 and passes through the dump heat exchanger DHE, exchanging heat with a brine solution connected to a source from which low-temperature heat can be collected, such as an external air collector, a solar collector, or a borehole drilled into the ground. Due to the heat exchange with the brine solution flowing from the large opening O4 to the large opening O3, the predominantly liquid refrigerant evaporates. The heat exchange between the brine solution and the refrigerant occurs under co-current conditions, which are known to have inferior heat exchange performance compared to counter-current heat exchange.

[0096] Just before leaving the dump heat exchanger DHE through the large opening O1, the refrigerant (now almost completely evaporated) exchanges heat with the relatively hot liquid refrigerant that enters the dump heat exchanger through the small opening SO2 and leaves the dump heat exchanger through the small port opening SO1. According to one embodiment of the present invention, when the refrigerant begins to exchange heat with the hot liquid refrigerant, approximately 85-98% (preferably 90-95% and more preferably 91-94%, for example 93%) of the refrigerant is evaporated.

[0097] Consequently, the temperature of the refrigerant that is about to leave the dump heat exchanger DHE through the opening O1 will increase, thereby ensuring that all of this refrigerant is fully evaporated.

[0098] Therefore, the low-temperature gaseous refrigerant entering the suction gas heat exchanger includes a certain amount of low-temperature liquid refrigerant, which is evaporated due to heat exchange with the high-temperature liquid refrigerant from the condenser. For example, the amount of low-temperature liquid refrigerant is 2-15 mass percent, preferably 5-10 mass percent, more preferably 6-9 mass percent, for example 7 mass percent.

[0099] It is well known to those skilled in the art that co-current heat exchange is inferior to counter-current heat exchange in terms of heat exchange performance. However, because heat exchange occurs between the relatively hot liquid brine entering the small-opening SO2 and the primarily gaseous refrigerant leaving the dump heat exchanger (DHE) (referred to as "suction gas heat exchange"), the refrigerant does not need to be fully evaporated during the brine-refrigerant heat exchange. Instead, the refrigerant may only be partially evaporated upon entering the suction gas heat exchange with the hot liquid refrigerant, as the remaining liquid refrigerant will evaporate during this heat exchange. Liquid-to-liquid heat exchange is generally more efficient than gas-to-liquid heat exchange. Co-current heat exchange has the additional benefit of reducing the risk of freezing, as the refrigerant enters the heat exchanger at a location with a high-temperature medium (with which it will exchange heat), thereby reducing the risk of freezing at this location, which is the most critical location for freezing.

[0100] Testing has shown that there may be issues with cold-starting the chiller system in cold environments.

[0101] From the opening O1 of the dump heat exchanger, the gaseous refrigerant will enter the four-way valve FWV, which is controlled to direct the flow of the gaseous refrigerant to the compressor, where the refrigerant is compressed again.

[0102] exist Figure 11a 、 Figure 11b In Figure 1, the chiller system is shown in cooling mode. To switch from heating to cooling mode, four-way valve FWV is controlled so that the compressor feeds compressed gaseous refrigerant to opening O1 of dump heat exchanger DHE. Expansion valve EXPV2 is fully closed, check valve OWV2 is opened, check valve OWV1 is closed, and expansion valve EXPV1 is opened, controlling the pressure before and after the refrigerant passes through expansion valve EXPV1.

[0103] Thus, in cooling mode, the dump heat exchanger will act as a countercurrent condenser and its "suction gas heat exchanger" will not perform any heat exchange, while the payload heat exchanger PLHE will act as a co-current evaporator. However, since suction gas heat exchange is provided between the hot liquid refrigerant and the semi-evaporated refrigerant about to leave the payload heat exchanger PLHE, the efficiency of the co-current heat exchange can be maintained at an acceptable level.

[0104] It should be noted that in Figure 10a 、 Figure 10b and Figure 11a 、 Figure 11b In the embodiment, the suction gas heat exchange portion is formed integrally with the dump heat exchanger DHE and the payload heat exchanger PLHE. However, in other embodiments, the suction gas heat exchanger may be separated from the dump heat exchanger and / or the payload heat exchanger.

[0105] In different climate zones, the demand for cooling and heating is different. In warmer climates, the demand for cooling is greater, where the refrigeration system will be used closer to full cooling effect and corresponding capacity in the suction gas heat exchanger is needed to evaporate any droplets that would otherwise leave the evaporator. For example, the evaporator is the payload heat exchanger PLHE in the cooling mode of the refrigeration system described above, where its integrated suction gas heat exchanger is used accordingly via a balancing valve BV that can be connected to the evaporator. Figure 11b The same or another balancing valve schematically shown in FIG. When the refrigeration system is used at reduced efficiency, for example at 25% or 50% of full efficiency, the suction gas heat exchanger is controlled by a balancing valve BV. The refrigeration system is reversible and can be switched between cooling mode and heating mode by means of a four-way valve FWV as described above. As shown, both the payload heat exchanger and the dump heat exchanger include integrated suction gas heat exchangers that can be activated and controlled by a balancing valve BV to ensure that the refrigerant evaporates before leaving the evaporator in both cooling mode and heating mode, and has zero superheat depending on the effect of the system operation. Thus, the amount of refrigerant directed to the suction gas heat exchanger can be adapted to the system conditions in both heating mode and cooling mode to provide an efficient reversible refrigeration system for different types of climates.

[0106] In another embodiment of the present invention, for example Figure 12 The heat exchanger shown, a "standard" heat exchanger 100, may be provided with a retrofit port heat exchanger 400 (see Figure 13 and Figure 14 ), which consists of some structure that fits into or just outside the port openings O1-O4 of a standard heat exchanger.

[0107] In the embodiment shown, the retrofit port heat exchanger 400 includes a tube 410 that fits within the port opening, the tube being bent in a semi-helical manner to allow high-temperature liquid refrigerant to flow therein in the same manner as the refrigerant flowing between the small port openings SO1 and SO2 of the previous embodiments, exchanging heat with the cold gaseous (or semi-gaseous) refrigerant about to leave the dump heat exchanger DHE or the payload heat exchanger PLHE.

[0108] Reference Figure 15 , schematically shows a cross section of a portion of a heat exchanger comprising a first heat exchanger plate 110 and a second heat exchanger plate 120 according to another embodiment. Figure 15In an embodiment, the first heat exchanger plate 110 is a symmetrical heat exchanger plate, wherein the second heat exchanger plate 120 is an asymmetrical heat exchanger plate as described above. Therefore, the corrugation depth of the first heat exchanger plate 110 is constant, while the corrugation depth of the second heat exchanger plate 120 is variable. The second heat exchanger plate 120 is formed with at least two different corrugation depths. Moreover, the first heat exchanger plate 110 and the second heat exchanger plate 120 are formed with corrugation patterns of different angles, such as the herringbone angle as described above. Figure 15 In the embodiment, the herringbone angle of the first heat exchanger plate 110 is 54 degrees, while the herringbone angle of the second heat exchanger plate 120 is 61 degrees. For example, the inter-plate volumes between adjacent plates are different, such that the inter-plate volume on one side of the first heat exchanger plate 110 is different from the inter-plate volume on the opposite side of the first heat exchanger plate 110. Of course, this also applies to the second heat exchanger plate 120. Therefore, the inter-plate volume between the first and second heat exchanger plates is different from the inter-plate volume between the second and first heat exchanger plates. Similarly, the cross-sectional area on one side of the first heat exchanger plate 110 is different from the cross-sectional area on the opposite side of the first heat exchanger plate 110.

[0109] Reference Figure 16 , schematically shows a cross section of a portion of a heat exchanger comprising a first heat exchanger plate 110 and a second heat exchanger plate 120 according to yet another embodiment. Figure 16 In the embodiment of , the first heat exchanger plate 110 is a symmetrical heat exchanger plate, and the second heat exchanger plate 120 is an asymmetrical heat exchanger plate as described above. Figure 16 In the embodiment of FIG. 5 , the herringbone angle of the first heat exchanger plate 110 is 45 degrees, while the herringbone angle of the second heat exchanger plate 120 is 61 degrees.

[0110] Reference Figure 17 , schematically shows a cross section of a portion of a heat exchanger comprising a first heat exchanger plate 110 and a second heat exchanger plate 120 according to yet another embodiment. Figure 17 In the embodiment of the present invention, the first heat exchanger plate 110 is an asymmetric heat exchanger plate, and the second heat exchanger plate 120 is also an asymmetric heat exchanger plate. Figure 17 In the embodiment, the herringbone angle of the first heat exchanger plate 110 is different from the herringbone angle of the second heat exchanger plate 120 as described above. Furthermore, the inter-plate flow channels have different volumes as described above. For example, the brazed joints are elongated, such as elliptical, and are arranged in a first orientation in the inter-plate flow channels with the larger volume, and in a different second orientation in the inter-plate flow channels with the smaller volume.

[0111] Reference Figure 18, schematically shows a cross section of a portion of a stack of first and second heat exchanger plates 110, 120 according to a further embodiment. Figure 18 In an embodiment, the first and second heat exchanger plates 110, 120 have different corrugation depths. The first heat exchanger plate 110 is a symmetrical heat exchanger plate, while the second heat exchanger plate 120 is an asymmetrical heat exchanger plate. Alternatively, both the first heat exchanger plate 110 and the second heat exchanger plate 120 are symmetrical or asymmetrical. The herringbone angle of the first heat exchanger plate 110 is different from the herringbone angle of the second heat exchanger plate 120, and the inter-plate flow channel volumes formed by the first and second heat exchanger plates 110, 120 when brazed together in a brazed joint are different.

[0112] Heat exchangers according to various embodiments of the present invention are used, for example, for condensation or evaporation, where at least one medium is in a gaseous phase at some point. For example, heat exchangers are used for heat exchange, where condensation or evaporation occurs in interplate flow channels with a relatively large volume. For example, a liquid medium (such as water or brine) is directed through interplate flow channels with a relatively small volume.

[0113] exist Figure 19 , an exemplary brazed true dual heat exchanger 500 comprising two separate, integrated suction gas heat exchangers ISGHX1 and ISGHX2 is shown in exploded view. True dual heat exchangers are used in heat pumps or chillers requiring high power ratios. Systems for true dual heat exchangers are well known to those skilled in the art and typically consist of two separate heat pump systems using true dual heat exchangers, rather than two separate heat exchangers.

[0114] The true dual heat exchanger 500 includes six heat exchanger plates 510, 520, 530, and 540. Each heat exchanger plate is provided with a pressed pattern of ridges and grooves, which are suitable for maintaining the plates at a certain distance from each other so that inter-plate flow channels 510-520, 520-530, 530-540, 540-510, 510-520 for heat exchange between the heat exchanger plates are formed. In addition, each heat exchanger plate is provided with port openings 550, 560, 570, 580, 590, 600, 610 for refrigerant and two port openings 620, 630 for water or brine solution. The port openings are selectively in fluid communication with the inter-plate flow channels in the following manner:

[0115] Port openings 630 and 640 are in fluid communication with interplate flow channels 510-520 and 530-540, port openings 550 and 560 are in fluid communication with interplate flow channels 520-530, port openings 570 and 580 are in fluid communication with interplate flow channels 540-510, and port openings 590, 600, 610, and 620 are in fluid communication with interplate flow channels 510-520.

[0116] The heat exchanger plates 510, 520, 530 and 540 are divided into subsections, in which the inter-plate flow channels are connected and restricted in certain ways: in the main section 650, all inter-plate flow sections are used for heat exchange with the medium; in a first isghx (integrated suction gas heat exchanger) section ISGHX1, the inter-plate flow channels 520-530 are fluidly connected to the inter-plate flow channels 520-530 of the main section, and one or two of the inter-plate flow channels 510-520 and / or 530-540 are connected to the port openings 610 and 620; in a second isghx section ISGHX2, the inter-plate flow channels 540-510 are fluidly connected to the inter-plate flow channels 540-510 of the main section, and one or two of the inter-plate flow channels 510, 520 and / or 530-540 are fluidly connected to the port openings 590, 600.

[0117] The main section is defined by the ISGHX sections ISGHX1 and ISGHX2 via a partition wall 660 that extends from one long side of each heat exchanger plate to the other. The partition wall comprises plate surfaces arranged at different heights, such that the mating of these plate surfaces of adjacent plates seals off the inter-plate flow channels 510-520 and 530-540 from the corresponding inter-plate flow channels of the ISGHX sections ISGHX1 and ISGHX2. Furthermore, the plate surfaces of the partition wall 660 are configured such that the mating of the plate surfaces of adjacent plates seals off the inter-plate flow channels 520-530 of the main section from the corresponding inter-plate flow channels of the second ISGHX section ISGHX2, and seals off the inter-plate flow channels 540-510 of the main section from the corresponding inter-plate flow channels of the first ISGHX section ISGHX1. The partition wall 660 separates the heat exchanger plates 510-540 into the main section 650 and the ISGHX sections ISGHX1 and ISGHX2. Thus, four port openings are provided in the main portion 650 , namely port openings 550 , 570 , 630 and 640 , while port openings 560 and 580 and first and second portions ISGHX1 and ISGHX2 having port openings 610 , 620 , 590 , 600 are provided on the other side of the partition wall 660 .

[0118] A second partition wall 670 is disposed between the ISGHX sections ISGHX1 and ISGHX2, extending from the short sides of the heat exchanger plates and the partition wall 660. The plate surfaces of this partition wall are arranged so that the plate surfaces of adjacent plates contact each other, thereby closing all inter-plate flow channels of the ISGHX sections ISGHX1 and ISGHX2 and preventing them from communicating with each other. Thus, the port opening 560 and the first ISGHX section having port openings 610 and 620 are arranged on one side of the partition wall 670, while the port opening 580 and the second ISGHX section having port openings 590 and 600 are arranged on the other side of the partition wall 670. Thus, the main portion 650m of the first ISGHX section ISGHX1 and the second ISGHX section ISGHX2 are separated by the partition walls 660 and 670.

[0119] Finally, each heat exchanger plate is provided with a skirt 680 extending around the entire perimeter of the heat exchanger plates 510, 520, 530, 540. The skirts 680 of adjacent plates are adapted to contact one another to form a circumferential seal, thereby preventing the medium from escaping from the inter-plate flow channels. Furthermore, the heat exchanger 500 according to the present invention is preferably provided with a start plate and / or end plate (not shown) arranged on either side of the stack of heat exchanger plates. In order to form a seal on the side of the port opening (which is not provided with a connection for the fluid exchanging heat to enter or leave the heat exchanger), one of the start plate and the end plate is provided with a port opening, while the other is not.

[0120] With the above arrangement, the true dual heat exchanger has separate interplate flow channels above the interplate flow channels 510-520 and 530-540 of the main portion 650, between the port openings 630 and 640; above the interplate flow channels 520-530 of the main portion and the first isghx portion ISGHX1, between the port openings 550 and 560; above the interplate flow channels 540-510 of the main portion 650 and the second isghx portion ISGHX2, between the port openings 570 and 580; above the interplate flow channels 520-530 of the first isghx portion ISGHX1, between the port openings 610 and 620; and above the interplate flow channels 540-510 of the second isghx portion ISGHX2, between the port openings 590 and 600.

[0121] Selective fluid communication between the port openings and the inter-plate flow channels can be achieved in a variety of ways, such as by providing surfaces around the port openings at different heights so that the surfaces of adjacent plates do or do not contact each other. Alternatively, selective fluid communication can be achieved by providing a separate sealing ring in the port opening, the sealing ring being provided with an opening for allowing communication where desired.

[0122] Furthermore, it should be noted that although described as a brazed heat exchanger, the true dual heat exchanger according to the present invention may be designed as a gasketed heat exchanger.

[0123] The true dual heat exchanger 500 according to the invention is particularly suitable for heat pump or chiller applications, where dual compressors are used in order to obtain a large ratio between low power and high power.

[0124] As above Figure 2-Figure 9 As described above, the heat exchanger plates 510-540 are provided with a first and second pattern of ridges R1, R2a, R2b and grooves G1, G2a, G2b. For example, every other heat exchanger plate is provided with the first pattern, while the other heat exchanger plates are provided with the second pattern. For example, heat exchanger plates 510 and 530 are provided with the first pattern, while heat exchanger plates 520 and 540 are provided with the second pattern, or vice versa. The first and second pressed patterns are, for example, herringbone patterns with different herringbone angles as described above or pressed patterns with diagonal lines at different angles, for example, with reference to Figure 6a 、 Figure 6b 、 Figure 7a and Figure 7b As described above. Main portion 650 has this pattern, for example, first and second isghx portions ISGHX1 and ISGHX2 also have this pattern. For example, the angle β1 (e.g., herringbone angle β1) of every other heat exchanger plate (e.g., heat exchanger plates 510, 530) is 25° to 70° or 30° to 45°. For example, the angle β2 (e.g., herringbone angle β2) of every other heat exchanger plate (e.g., heat exchanger plates 520, 540) is 25° to 70° or 30° to 45°. The first and second patterns are in opposite directions, such that the angles or herringbone vertices alternate in opposite directions throughout the heat exchanger. For example, the difference between the first herringbone angle β1 and the second herringbone angle β2 is 2° to 35°.

[0125] For example, the grooves G1 of every other heat exchanger plate are formed to have the same depth D1, as shown in FIG. Figure 3 As described above, the other heat exchanger plates having the first and second grooves G2a, G2b are formed to have different depths, wherein the first groove G2a is formed to have a first depth D2a and the second groove G2b is formed to have a second depth D2b, as shown in FIG. Figure 4 Thus, every other inter-plate flow channel has a larger volume than the rest, also as described above.

[0126] For example, as referenced Figure 8 and Figure 9As described, the contact points and the brazing joints are arranged alternately so that the brazing joints between the heat exchanger plates 510-540 are elongated, for example elliptical, wherein the brazing joints are arranged in a first direction in the inter-plate flow channels with a larger volume and in a second direction in the inter-plate flow channels with a smaller volume.

[0127] Reference Figure 20 , schematically shows a first pattern of ridges R1 and grooves G1 of a first heat exchanger plate 110. Figure 20 In the embodiment, the first heat exchanger plate 110 includes small port openings SO1, SO2 and a dividing surface DW to provide a first heat exchange portion 130 and a second heat exchange portion 140 forming an integrated suction gas heat exchanger as described above. Optionally, the first heat exchanger plate 110 includes dividing walls 660, 670 and small port openings 590-620 to provide Figure 19 Two integrated suction gas heat exchangers ISGHX1, ISGHX2 are shown. Figure 20 The pressing pattern of the embodiment is a herringbone pattern, but may alternatively be a diagonal pattern, thus having a substantially Figure 6a and Figure 6b The first angle β1 shown is not included, but is located in the central main heat exchange portion of the heat exchanger plate 110. Therefore, the first embossed pattern partially includes the first angle β1. For example, the central main heat exchange portion extends across the first heat exchanger plate 110 from one side to the opposite side. The central main heat exchange portion is arranged between the first and second heat exchange portions at the port openings of the heat exchanger plate, referred to herein as end portions. The first and second end portions are, for example, arranged at opposite ends of the first heat exchanger plate 110. For example, the first and second end portions extend through the first heat exchanger plate 110 from one side to the opposite side. The first end portion includes port openings, such as the first and third port openings O1 and O3, as well as small port openings SO1 and SO2, and a dividing surface DW, forming an extraction gas heat exchanger. The second end portion includes port openings, such as the second and fourth port openings O2 and O4. In at least one end portion, such as the first and second end portions, the embossed pattern of ridges and grooves R1 and G1 is arranged at an angle β1', which differs from the angle β1 of the embossed pattern in the central main heat exchange portion. For example, the direction of the pressed pattern in the central main portion is the same as the direction of the pressed pattern in the end portions. For example, the angles are the same in both end portions. Alternatively, the angle of the first end portion is different from the angle of the second end portion. Optionally, the second heat exchange portion 140 is arranged in a different pattern or angle than the first end portion. Figure 20, the first heat exchanger plate 110 is shown as an example, but it is understood that the second pressed pattern of the second heat exchanger plate 120 is designed in a corresponding manner, wherein the second pattern of ridges R2a, R2b and grooves G2a, G2b is arranged at an angle β2 in the central main heat exchange portion and at a different angle β2' (not shown) in the end portions.

Claims

1. A refrigeration method comprising the following steps: In the compressor, the gaseous refrigerant is compressed, causing the temperature and pressure of the gaseous refrigerant to increase, wherein the boiling point of the gaseous refrigerant is increased; supplying the gaseous refrigerant to a condenser, wherein the gaseous refrigerant from the compressor exchanges heat with a high-temperature heat carrier, the heat exchange causing the refrigerant to condense; reducing the pressure of the liquid refrigerant from the condenser through an expansion valve, thereby lowering the boiling point of the refrigerant; supplying the low-boiling-point refrigerant to an evaporator, wherein the low-boiling-point refrigerant exchanges heat with a low-temperature heat carrier in a downstream flow manner in the evaporator, so that the refrigerant is partially evaporated; as well as, In a suction gas heat exchanger integrated with the evaporator, heat is exchanged between the high-temperature liquid refrigerant from the condenser and the low-temperature gaseous refrigerant from the evaporator. The low-temperature gaseous refrigerant entering the suction gas heat exchanger contains a certain amount of low-temperature liquid refrigerant of 6-9 mass %, and the low-temperature liquid refrigerant evaporates by heat exchange with the high-temperature liquid refrigerant from the condenser. Wherein, the refrigeration method further comprises: measuring the temperature of the high-temperature liquid refrigerant, measuring the temperature of the low-temperature gaseous refrigerant, calculating a temperature difference between the high-temperature liquid refrigerant and the low-temperature gaseous refrigerant, and controlling a balancing valve to bypass the suction gas heat exchanger when the temperature difference between the high-temperature liquid refrigerant and the low-temperature gaseous refrigerant is less than a predetermined threshold, The refrigeration method further includes the step of reversing the flow of the refrigerant through a four-way valve to switch between the cooling mode and the heating mode. The condenser and evaporator are respectively brazed plate heat exchangers (100) comprising first and second heat exchanger plates (110, 120), wherein the first and second heat exchanger plates (110, 120) have intersecting ridges and grooves connected by elongated brazed joints (170) and form inter-plate flow channels of different sizes.

2. The refrigeration method according to claim 1, wherein: The amount of the certain amount of low-temperature liquid refrigerant is 7% by mass.

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