KR20170085936A - Refrigerator for super-freezing a storing chamber - Google Patents

Abstract

An embodiment of the present invention relates to a deep temperature freezer.
The deep-sea freezer according to the embodiment of the present invention includes a plurality of heat exchangers installed in the suction pipe and performing heat exchange of the mixed refrigerant sucked into the compressor.

Description

[0001] The present invention relates to a refrigerator for super-

An embodiment of the present invention relates to a deep temperature freezer.

A deep-sea freezer is understood as a device for driving a refrigeration cycle to form a storage room at a temperature of -60 ° C to -80 ° C or less in a cryogenic environment.

To realize the refrigeration cycle, a refrigerant having a low boiling point can be used. However, when the refrigerant having a low boiling point is used alone, there is a problem that the discharge pressure of the compressor rises and the reliability of the compressor deteriorates.

Therefore, two or more mixed refrigerants having different boiling points may be used in the refrigeration cycle for realizing the deep temperature. The mixed refrigerant contains a mixed refrigerant that does not change its temperature in a quasi-equilibrium state when the liquid is vaporized between a liquid phase and a vapor phase under a predetermined pressure, that is, an azeotropic refrigerant mixture, A non-azeotropic refrigerant mixture is included.

The azeotropic mixed refrigerant exists only in a specific component ratio and exhibits thermodynamic properties such as pure water. On the other hand, the non-azeotropic mixed refrigerant may vary in the evaporation pressure or temperature depending on its composition

On the other hand, since the azeotropic mixed refrigerant has a disadvantage that it is difficult to realize a deep temperature (cryogenic temperature), the non-azeotropic mixed refrigerant may be preferably used to realize the deep temperature.

However, even if the non-azeotropic mixed refrigerant is used, the discharge pressure or the condensation pressure of the compressor is relatively high. Therefore, it is necessary to select a compressor suited to the discharge pressure range. In general, a compressor used in a deep temperature freezer may be a commercial compressor having a large operating pressure range, that is, a high discharge pressure value.

However, the commercial compressor has a problem that operation noise of a deep temperature freezer is deteriorated due to a large operation noise.

Meanwhile, the information on the prior art related to the deep freezer is as follows.

1. Registration number (Registration date): US 7,299,653B2 (November 27, 2007)

2. Description of the Invention: The refrigeration system using non-azeotropic refrigerant, and non-azeotropic refrigerant for very low temperature used for the system.

According to the above-mentioned prior art, although a deep temperature environment can be realized by using two or more mixed refrigerants, there is a problem that it is difficult to satisfy both the deep temperature implementation and the proper discharge pressure of the compressor because the mixing ratio thereof is not optimized.

In detail, when the ratio of the refrigerant having a high boiling point is high, it is difficult to realize the deep temperature, and when the ratio of the refrigerant having a low boiling point is high, the discharge pressure of the compressor is increased and the reliability of the compressor is problematic.

Meanwhile, in the refrigeration cycle disclosed in the above prior art, the refrigerant discharged from the compressor is condensed in the condenser, performs heat exchange with the evaporative refrigerant, and is configured to realize the deep temperature by the heat exchange.

However, according to the above-mentioned prior art document, there is a problem that the quality of the refrigerant increases in the process of being expanded in the expansion apparatus after the heat exchange, and accordingly, the ratio of the liquid refrigerant in the refrigerant flowing into the evaporator is reduced, thereby reducing the cooling power.

The present invention has been made in order to solve the above problems, and it is an object of the present invention to provide a deep-sea freezer capable of realizing a desired cryogenic environment.

It is another object of the present invention to provide a deep-sea freezer capable of lowering the condensation pressure of a refrigeration cycle.

It is another object of the present invention to provide a deep-sea freezer that can reduce noise generated in a compressor and thereby increase the reliability of the compressor.

The deep-sea freezer according to the embodiment of the present invention includes a plurality of heat exchangers installed in the suction pipe and performing heat exchange of the mixed refrigerant sucked into the compressor.

Wherein the plurality of heat exchangers include a first heat exchanger, and the first heat exchanger includes: a first suction heat exchanger for guiding the flow of the mixed refrigerant sucked into the compressor; And a condensation heat exchanger for performing heat exchange with the first suction heat exchanger and guiding the flow of the condensation pipe.

The length of the first suction heat exchanging or condensing heat exchanging part is formed within a range of 3.5 to 5 m.

The diameter of the pipe of the condensation heat exchanger is larger than the diameter of the pipe of the expansion device.

The pipe diameter of the condensation heat exchanger is formed within a range of 3.5 to 4.5 times the pipe diameter of the expansion device.

The second heat exchanger includes a second heat exchanger and the second heat exchanger includes a second suction heat exchanger provided at one side of the first suction heat exchanger and guiding the flow of mixed refrigerant sucked into the compressor; And the expansion device performing heat exchange with the second suction heat exchanger.

Wherein the first suction heat exchanger and the condensing pipe or the second suction heat exchanger and the expansion device are in contact with each other to perform heat exchange.

The first suction heat exchanger and the condensing pipe or the second suction heat exchanger and the expansion device are coupled by soldering.

And a heat exchanger connecting pipe disposed between the first and second heat exchangers for preventing heat exchange between the expansion device and the condensation heat exchanging part, wherein the first heat exchanging part and the second heat exchanging part are connected to the heat exchanger And are separated from each other by a connection pipe.

The first heat exchanger is installed on the outlet side of the second heat exchanger on the basis of the flow direction of the refrigerant flowing through the suction pipe.

The second heat exchanger is installed on the outlet side of the first heat exchanger on the basis of the flow direction of the refrigerant flowing through the condensing pipe.

The evaporator includes a first evaporator and a second evaporator connected in series to each other, and the second evaporator is installed at an outlet side of the first evaporator.

Wherein the evaporator includes a first evaporator and a second evaporator connected in parallel to each other, wherein the expansion device includes a first expansion device installed at an inlet side of the first evaporator and a second expansion device installed at an inlet side of the second evaporator 2 expansion device.

The first and second expansion devices and the suction pipe are coupled to each other and heat-exchanged.

Wherein the evaporator includes: a first evaporator installed at an outlet side of the expansion device; A second evaporator connected in series to an outlet side of the first evaporator; And a third evaporator connected in series to the outlet side of the second evaporator.

Two independent refrigeration cycles are driven, and each of the independent refrigeration cycles includes the compressor, the condenser, the expansion device, the evaporator, and a plurality of heat exchangers.

The mixed refrigerant includes butane and ethylene.

According to the embodiment of the present invention, the refrigerant condensed in the condenser is allowed to pass through the plurality of heat exchangers before being introduced into the evaporator, thereby lowering the condensation pressure of the refrigeration cycle and preventing the rise of the condensed refrigerant as it passes through the expansion device The effect can be seen.

In more detail, since the plurality of heat exchangers include a first heat exchanger that performs heat exchange between the refrigerant that has passed through the condenser and the suction refrigerant that is sucked into the compressor, the condensation pressure of the refrigeration cycle in the heat exchange process in the first heat exchanger It can be lowered. As a result, there is an advantage that it is possible to use a household compressor having low discharge pressure and low noise, which is used in a general refrigerator.

Further, since the temperature of the suction refrigerant rises, the liquid refrigerant can be prevented from flowing into the compressor, thereby improving the operational reliability of the compressor.

The plurality of heat exchangers include a second heat exchanger for exchanging heat between the refrigerant passing through the expansion device after heat exchange in the first heat exchanger and the suction refrigerant sucked into the compressor, It is possible to prevent an increase in the degree of dryness in the course of decompression.

As a result, the ratio of the liquid refrigerant in the refrigerant flowing into the evaporator increases, and the evaporative heat amount, that is, the cooling power, can be improved.

Since the tube diameter of the condensing heat exchanger constituting the first heat exchanger is larger than the tube diameter of the expansion device constituting the second heat exchanger, condensation can be easily performed in the process of passing the refrigerant through the first heat exchanger, The condensation temperature and the condensation pressure can be lowered.

Also, since the length of the first heat exchanger is suggested to be the optimum range and heat exchange can be performed, the operation condition of the domestic compressor can be satisfied and the operation reliability of the compressor can be improved due to the characteristics of the refrigerant cycle.

The first heat exchanger may be formed by a combination of a condensing pipe and a suction pipe, and the second heat exchanger may be formed by a combination of a capillary tube and a suction pipe, thereby improving heat exchange efficiency.

In addition, since the condensed refrigerant passing through the condenser can be introduced into the second heat exchanger after heat-exchanged in the first heat exchanger, the condensation pressure can be lowered first, and the rise in the degree of evaporation during the expansion process can be prevented.

In addition, since the weight ratio of the non-azeotropic mixed refrigerant can be optimally suggested, the desired cryogenic temperature environment can be achieved and the appropriate discharge pressure of the domestic compressor can be satisfied.

1 is a view showing a refrigeration cycle of a deep-sea freezer according to a first embodiment of the present invention.
FIG. 2 is a view showing a configuration of a first and a second heat exchanger according to a first embodiment of the present invention.
3 is a ph diagram of the deep-sea freezer according to the first embodiment of the present invention.
4 is an experimental graph showing the optimal range of the length of the first heat exchanger according to the first embodiment of the present invention.
5 is an experimental graph showing a plurality of resultant values varying according to the refrigerant amount of the non-azeotropic mixed refrigerant according to the first embodiment of the present invention.
6 is a view illustrating a refrigeration cycle of the deep-sea freezer according to the second embodiment of the present invention.
7 is a view showing a refrigeration cycle of the deep-sea freezer according to the third embodiment of the present invention.
FIG. 8 is a view showing a refrigeration cycle of the deep-sea freezer according to the fourth embodiment of the present invention.
9 is a view showing a refrigeration cycle of the deep-sea freezer according to the fifth embodiment of the present invention.

Hereinafter, specific embodiments of the present invention will be described with reference to the drawings. It is to be understood, however, that the spirit of the invention is not limited to the embodiments shown and that those skilled in the art, upon reading and understanding the spirit of the invention, may easily suggest other embodiments within the scope of the same concept.

FIG. 1 is a view showing a refrigeration cycle of a deep-sea refrigerating machine according to a first embodiment of the present invention, FIG. 2 is a view showing the construction of a first and a second heat exchanger according to a first embodiment of the present invention, 3 is a ph diagram of the deep-sea freezer according to the first embodiment of the present invention, FIG. 4 is an experimental graph showing an optimal range of the length of the first heat exchanger according to the first embodiment of the present invention, and FIG. Is an experimental graph showing a plurality of resultant values varying according to the refrigerant amount of the non-azeotropic refrigerant according to the first embodiment of the present invention.

Referring to FIG. 1, a refrigeration cycle in which refrigerant compression, condensation, expansion and evaporation are repeated can be operated in the deep-sea freezer 10 according to the first embodiment of the present invention. And a compressor 110 capable of compressing the refrigerant. The compressor 110 may include a household compressor used in a general household refrigerator.

For example, the temperature or pressure operating range of the compressor 110 is as follows. The compressor 110 may be configured to have a maximum discharge pressure of 25 bar or less, a maximum discharge temperature of 120 ° C or less, and a minimum suction pressure of 1 bar or less. The domestic compressor (110) having such a temperature or pressure range has an advantage that very little operation noise is generated.

The refrigerant sucked into the compressor (110) includes mixed refrigerant. The mixed refrigerant includes a first refrigerant having a first boiling point and a second refrigerant having a second boiling point lower than the first boiling point. The first refrigerant may be referred to as a " high-temperature refrigerant ", and the second refrigerant may be referred to as a "low-temperature refrigerant ".

As the mixed refrigerant is included in the refrigerant, it is possible to realize the evaporation temperature required in the deep-sea freezer, that is, the deep temperature, and the pressure of the refrigerant discharged from the compressor 110 can be set within the set range.

In detail, it is possible to realize a deep temperature by the characteristics of the low-temperature refrigerant. However, when the low-temperature refrigerant is compressed by the compressor 110, the refrigerant has a relatively high discharge pressure, thereby adversely affecting the reliability of the compressor, particularly, the domestic compressor 110 applied to the present embodiment. Therefore, in order to lower the discharge pressure, a high-temperature refrigerant having a relatively low discharge pressure may be mixed.

However, merely mixing the low-temperature refrigerant and the high-temperature refrigerant may cause the discharge pressure of the mixed refrigerant to be higher than the operating pressure of the domestic compressor 110 used in the present embodiment.

In order to solve such a problem, a commercial compressor having a large operating pressure range can be used, but in this case, the reliability of the deep temperature freezer may be lowered due to a very large operation noise. Therefore, in the present embodiment, the ratio of the mixed refrigerant that can meet the operating pressure or operating temperature range of the domestic compressor 110, that is, the proper ratio of the high-temperature refrigerant and the low-temperature refrigerant, is proposed.

Generally, the high-temperature refrigerant may include isopentane, 1,2-butadiene, N-butane, 1-butene or isobutane, . The physical properties of the high-temperature refrigerant are shown in Table 1 below.

High temperature refrigerant Evaporation temperature (1 bar), ℃ Evaporation temperature (20 bar), ℃ ISOPENTANE 27.5 154.7 1,2-BUTADIENE 10.3 124.8 N-BUTANE -0.9 114.5 1-BUTENE -6.6 105.8 ISOBUTANE -12 100.7

In the case of isopentane and 1,2-butadiene based on the pressure of 1 bar, the evaporation temperature tends to be somewhat higher. Therefore, when the isopentane and 1,2-butadiene are used as the high-temperature refrigerant according to the present embodiment, it is difficult to realize the deep temperature even if they are mixed with the low-temperature refrigerant Lt; / RTI >

On the other hand, in the case of N-butane, 1-butene and isobutane, the evaporation temperature is 0 ° C or lower, based on 1 bar. Therefore, any one of N-butane, 1-butene and isobutane can be used as a refrigerant for high temperature according to the present embodiment in a broad sense.

However, in the case of 1-butene (1-butene) and isobutane, the evaporation temperature is somewhat low, and when mixed with the low-temperature refrigerant, deep temperature can be realized. However, Problems may arise. Therefore, preferably, the high-temperature refrigerant according to the present embodiment is characterized by using N-butane whose evaporation temperature is close to 0 ° C on the basis of 1 bar.

Meanwhile, the low-temperature refrigerant may include ethane or ethylene. The physical properties of the low-temperature refrigerant are shown in Table 2 below.

Low temperature refrigerant Evaporation temperature (1 bar), ℃ Evaporation temperature (20 bar), ℃ ETHANE -88.8 -182.8 ETHYLENE -104 -169.15

On the basis of the pressure of 1 bar, in the case of the above-mentioned ethane, the evaporation temperature tends to be somewhat higher. Therefore, when the ethane is used as the low-temperature refrigerant according to the present embodiment, it is difficult to realize the deep temperature.

On the other hand, based on the pressure of 1 bar, in the case of ethylene, the evaporation temperature has a value of -100 ° C. or less, and the evaporation temperature of -100 ° C. or less forms a temperature suitable for realizing the deep temperature. Therefore, preferably, the low-temperature refrigerant according to the present embodiment is characterized in that ethylene having an evaporation temperature of -100 ° C or less based on 1 bar is used.

As described above, in order to meet the discharge pressure range of the domestic compressor 110, even when the N-butane is used as the high-temperature refrigerant and ethylene is used as the low-temperature refrigerant, , And they need to be mixed at an appropriate weight ratio.

In this embodiment, a number of experiments are repeated to propose the ratio of the mixed refrigerant that satisfies the discharge pressure range of the domestic compressor 110.

For example, the N-butane may be determined in a range of 80 wt% to 85 wt%, and the ethylene may be determined in a range of 15 wt% to 20 wt%.

In detail, we present the results for repeated experiments.

Room temperature (32 ℃) N-BUTANE / ETHYLENE (% by weight) 70/30 75/25 80/20 85/15 90/10 Maximum discharge pressure (bar) 39.1 36.2 24.3 22.9 21.4 Minimum suction pressure (bar) 1.7 1.5 1.18 1.09 0.94 Maximum discharge temperature (℃) 116.9 111.2 105.4 101.3 98.6 Temperature performance (℃) -79.8 -74.6 -68.3 -62.9 -55.8

Table 3 shows the results of experiments in which the weight percentages of butane (N-butane) and ethylene are different in the ambient temperature (room temperature) of 32 ° C.

Analysis of the above results shows that when the weight percent of butane is relatively increased in the mixed refrigerant of N-butane and ethylene, the maximum discharge pressure, minimum suction pressure and maximum discharge temperature of the compressor And the temperature performance, i.e., the temperature value of the storage room implemented in the deep-sea freezer, is increased.

As described above, in order to satisfy the operating pressure and temperature range of the domestic compressor 110, the maximum discharge pressure should be 25 bar or less, the maximum discharge temperature should be 120 ° C or less, and the minimum suction pressure should be 1 bar or more.

The ratio of N-butane to ethylene satisfying these conditions forms a weight percentage value ratio of 80:20 to 85:15. Within the range of this weight percentage, the temperature of the storage chamber capable of exhibiting the required performance of the deep temperature freezer, for example, a value of -60 DEG C or less can be formed.

Room temperature (38 ℃) N-BUTANE / ETHYLENE (% by weight) 70/30 75/25 80/20 85/15 90/10 Maximum discharge pressure (bar) 40.8 38.3 24.8 23.6 22.2 Minimum suction pressure (bar) 1.9 1.6 1.25 1.14 0.98 Maximum discharge temperature (℃) 121.3 118.4 108.3 103.6 101.8 Temperature performance (℃) -76.5 -72.4 -66.7 -61.9 -53.2

Table 4 shows the results of experiments in which the weight percentages of butane (N-butane) and ethylene were varied at an ambient temperature (room temperature) of 38 ° C.

Analysis of the above results shows that when the weight percent of butane is relatively increased in the mixed refrigerant of N-butane and ethylene, the maximum discharge pressure, minimum suction pressure and maximum discharge temperature of the compressor And the temperature performance, i.e., the temperature value of the storage room implemented in the deep-sea freezer, is increased.

As described above, in order to satisfy the operating pressure and temperature range of the domestic compressor 110, the maximum discharge pressure should be 25 bar or less, the maximum discharge temperature should be 120 ° C or less, and the minimum suction pressure should be 1 bar or more.

The ratio of N-butane to ethylene satisfying these conditions forms a weight percentage value ratio of 80:20 to 85:15. Within the range of this weight percentage, the temperature of the storage chamber capable of exhibiting the required performance of the deep temperature freezer, for example, a value of -60 DEG C or less can be formed.

To summarize, in order to achieve a desired performance of the deep-freezer 10 employing the domestic compressor 110, the ratio of N-butane to ethylene is 80:20 to 85:15 % Ratio value of the refrigerant can be used.

5 shows a result of performing an experiment in which a mixed refrigerant is formed such that the ratio of N-butane to ethylene is 83: 17, and the storage chamber having a predetermined volume is cooled while the amount of refrigerant is increased .

In detail, it is understood that it is a graph showing the temperature of the mixed refrigerant sucked into the compressor 110 as the amount of the refrigerant increases, that is, the temperature of the suction pipe, the temperature of the storage room to be cooled, and the amount of energy consumed by operation of the deep-

As the amount of the mixed refrigerant increases, the temperature of the intake pipe tends to decrease gradually, the temperature of the storage chamber also decreases, and the amount of consumed energy increases gradually.

According to FIG. 5, in order to have a desired temperature performance of a deep-sea refrigerating machine, that is, a value of -60 ° C or less, the amount of mixed refrigerant needs to be 80 g or more. Of course, depending on the volume of the storage chamber, the amount of the mixed refrigerant may be varied.

Referring to FIGS. 1 and 3 together, changes in the configuration and physical properties of the cycle will be described. 3 shows a Ph diagram, and a portion indicated by a dotted line represents a refrigeration cycle in which a plurality of heat exchangers 210 and 250 according to the present embodiment are not provided, and a portion indicated by a solid line represents a configuration according to the present embodiment In the refrigeration cycle.

In detail, a plurality of isotherms are displayed on the P-h line. The isotherms include T2 (T2 '), T3, T4, T5, and T7. The temperature value according to the isotherm can satisfy the following relationship: T2 (T2 ')> T7> T3> T5> T4. For example, T2 (T2 ') may be formed at 35 to 40 ° C, T7 at 30 to 35 ° C, T3 at 8 to 13 ° C, T5 at about -60 ° C, and T4 at about -80 ° C.

The deep temperature freezer 10 according to the first embodiment of the present invention further includes a condenser 120 installed at the outlet side of the compressor 110 for condensing the mixed refrigerant discharged from the compressor 110 . The deep-sea freezer 10 includes a dryer 130 installed at the outlet of the condenser 120 for filtering moisture or foreign matter from the refrigerant condensed in the condenser 120.

The deep-sea freezer (10) further includes an expansion device (140) installed at an outlet side of the dryer (130) and depressurizing the refrigerant condensed in the condenser (120). For example, the expansion device 140 may include a capillary tube.

The deep-sea freezer (10) further includes a condensing pipe (161) extending from the outlet side of the condenser (120) to the expansion device (140). The dryer 130 may be installed in the condensing pipe 161.

The deep-sea freezer 10 further includes an evaporator 150 installed at an outlet side of the expansion device 140 and evaporating the refrigerant decompressed in the expansion device 140. The deep heat freezer 10 further includes a suction pipe 165 extending from an outlet side of the evaporator 150 to a suction side of the compressor 110.

The refrigerant may be supplied to the storage room of the deep-sea freezer 10 through the refrigerant passing through the compressor 110, the condenser 120, the expansion device 140, and the evaporator 150.

The deep heat freezer (10) further includes a plurality of heat exchangers (210, 250) for improving the operation efficiency of the deep heat freezer (10).

The plurality of heat exchangers 210 and 250 include a first heat exchanger 210 for exchanging heat between a refrigerant flowing through the condensing pipe 161 and a refrigerant flowing through the suction pipe 165.

In detail, the first heat exchanger 210 may include a first suction heat exchanger 211 and a condensation heat exchanger 213 for performing heat exchange with the first suction heat exchanger 211. The first suction heat exchanger 211 constitutes at least a part of the suction pipe 165 and the condensation heat exchanger 213 may constitute at least a part of the condensation pipe 161.

The first suction heat exchanger 211 and the condensation heat exchanger 213 may be configured to contact each other. For example, the first suction heat exchanging part 211 and the condensing heat exchanging part 213 may be coupled by sodering.

When the heat exchange is performed between the first suction heat exchanger 211 and the condensing heat exchanger 213, the low-temperature refrigerant flowing through the first suction heat exchanger 211 flows into the condensing heat exchanger 213 at a high temperature The refrigerant can be cooled.

Accordingly, the condensing pressure of the refrigeration cycle is lowered, and thus the discharge pressure of the compressor 110 can be reduced. As the discharge pressure of the compressor 110 is reduced, the operation reliability of the domestic compressor 110 can be improved and the noise can be reduced as described above.

Since the heat absorbed by the refrigerant flowing through the first suction heat exchanger 211 can be made, the ratio of the liquid refrigerant included in the refrigerant, that is, the non-effective cooling power can be reduced. In addition, liquid refrigerant can be prevented from flowing into the compressor (110).

3, the state of the refrigerant (point 1) compressed by the compressor 110 represents point 2 after passing through the condenser 120. In FIG. Then, the refrigerant is condensed while passing through the condensation heat exchanger 213 (heat quantity Q1), and as a result, the condensation temperature is lowered from T2 to T3 and the condensation pressure is formed of Pd.

On the other hand, in the prior art, the refrigerant compressed in the compressor is condensed only in the condenser. At this time, the state of the compressed refrigerant represents point 1 ', and the state of the refrigerant after passing through the condenser represents point 2'. That is, as compared with the present invention, the condensation pressure indicates Pd 'higher than Pd and the condensation temperature indicates T2' higher than T3. Here, T2 'has the same temperature value as T2.

As a result, the refrigerant passing through the condenser 120 is heat-exchanged in the first heat exchanger 210 so that the condensation pressure Pd is lower by ΔP than the condensation pressure beam Pd ' ) Is lower than the condensation temperature T2 'of the prior art.

The refrigerant passing through the first suction heat exchanging part 211 may undergo heat absorption from the refrigerant passing through the condensing heat exchanging part 213 to evaporate (point 6 -> 7, heat quantity Q1 ').

The non-effective cooling power is understood to be a cooling power of a coolant having a temperature higher than the temperature of the cool air to be supplied to the deep-room storage chamber, for example, -60 ° C or more, which is difficult to supply to the deep-temperature storage chamber. That is, in the P-h line, since the temperature of T5 is about -60 DEG C, it can be understood that the cooling power of the refrigerant forming the temperature from T5 to T7 is a non-effective cooling power.

As a result, some of the non-effective cooling power may be used to cool the condensation heat exchanger 213 of the first heat exchanger 210 through the first suction heat exchanger 211. In this process, the specific gravity of the liquid refrigerant may decrease while the refrigerant passing through the first suction heat exchanging part 211 evaporates.

The plurality of heat exchangers 210 and 250 includes a second heat exchanger 250 for exchanging heat between a refrigerant flowing through the expansion device 140 and a refrigerant flowing through the suction pipe 165.

The first heat exchanger 210 may be installed at the outlet side of the second heat exchanger 250 with reference to the flow direction of the refrigerant flowing through the suction pipe 165. The second heat exchanger 250 may be installed on the outlet side of the first heat exchanger 210 on the basis of the flow direction of the refrigerant flowing through the condensing pipe 161.

In detail, the second heat exchanger 250 is provided with a second suction heat exchanger 251 provided at one side of the first suction heat exchanger 211 and a second suction heat exchanger 251 provided at one side of the first suction heat exchanger 211, Apparatus 140 may be included. The second suction heat exchanging part 251 may constitute at least a part of the suction pipe 165.

The second suction heat exchanger 251 and the expansion device 140 may be configured to contact each other. For example, the second suction heat exchanging part 251 and the expansion device 140 may be coupled by sodering.

When the heat exchange is performed between the second suction heat exchanger 251 and the expansion device 140, the low-temperature refrigerant flowing through the second suction heat exchanger 251 flows into the expansion device 140 through the high- Can be cooled. Accordingly, the refrigerant is depressurized in the process of passing through the expansion device 140, and the tendency of the process of increasing the dryness in the depressurization process can be reduced.

That is, when the refrigerant passes through the expansion device 140, the pressure and the temperature of the refrigerant are lowered, and the ratio of the gaseous refrigerant in the refrigerant can be increased. The gaseous refrigerant has an adverse effect on the evaporation performance of the evaporator 150. When the ratio of the gaseous refrigerant increases, the proportion of the liquid refrigerant that can evaporate is decreased, so that the evaporation performance may be deteriorated.

As a result, since the refrigerant passing through the expansion device 140 can be cooled through the second heat exchanger 250, the liquid refrigerant ratio at the inlet side of the evaporator 150 can be increased, .

The first heat exchanger 210 may be installed on the outlet side of the second heat exchanger 250 with respect to the refrigerant flow direction in the suction pipe 165. In other words, the first suction heat exchanging part 211 may be installed at the outlet side of the second suction heat exchanging part 251.

Therefore, the refrigerant having passed through the evaporator 150 is heat-exchanged in the second heat exchanger 250 and then heat-exchanged in the first heat exchanger 210, so that the dryness and the non-effective cooling power can be reduced. Then, the suction temperature can be increased in accordance with the increase in the dryness level, and the suction temperature condition of the domestic compressor 110 can be easily satisfied as the suction temperature rises.

3, the refrigerant passing through the first heat exchanger 210 passes through the expansion device 140 and is heat-exchanged with the second suction heat exchanger 251 (heat quantity Q2). As a result, The temperature of the refrigerant decreases from T3 to T4 and the pressure of the refrigerant can be lowered from Pd to Ps (the refrigerant state shifts from point 3 to point 4). As a result, the refrigerant heat-exchanged in the first heat exchanger 210 is further heat-exchanged in the second heat exchanger 250, so that the pressure and the temperature of the refrigerant can be lowered.

The refrigerant passing through the expansion device 140 of the second heat exchanger 250 flows into the evaporator 150 and evaporates. After the refrigerant passes through the evaporator 150, the state changes from point 4 to 5. The temperature T4 at the point 4 is about -80 DEG C, and the temperature T5 at the point 5 is about -60 DEG C. Therefore, the refrigerant cooling force in the section from point 4 to 5 can serve as an effective cooling power sufficient to cool the cold air to be supplied to the storage room of the deep-sea freezer.

The refrigerant that has passed through the evaporator 150 is absorbed from the refrigerant passing through the expansion device 140 and evaporates through the second suction heat exchanger 251 of the second heat exchanger 250 -> 6, calories Q2 '). The refrigerant cooling force in the section from point 5 to 6 acts as a non-effective cooling power, corresponding to a temperature of -60 ° C or higher.

However, the heat amount Q2 'may be used to cool the expansion device 140 of the second heat exchanger 250 through the second suction heat exchanger 251. [ In this process, the specific gravity of the liquid refrigerant may decrease while the refrigerant passing through the second suction heat exchanging part 251 evaporates.

In summary, Q1 'and Q2' are used to cool the non-effective cooling power or the expansion device 140 of the condensing heat exchanger 213 of the first heat exchanger 210 and the second heat exchanger 250 . In this process, the specific gravity of the liquid refrigerant may decrease while the refrigerant passing through the first and second suction heat exchangers (211, 251) evaporates.

The deep-sea freezer (10) further includes a heat exchanger connecting pipe (260) disposed between the first heat exchanger (210) and the second heat exchanger (250). The heat exchanger connecting pipe 260 constitutes a part of the condensing pipe 161 and may be configured to connect the first heat exchanger 210 and the second heat exchanger 250.

The first heat exchanger 210 and the second heat exchanger 250 are spaced apart from each other by the heat exchanger connecting pipe 260 to prevent heat exchange between the first and second heat exchangers 210 and 250 . That is, heat exchange between the condensing heat exchanger 213 and the expansion device 140 can be prevented.

If heat exchange is performed between the condensing heat exchanger 213 and the expansion device 140, the cooling effect of the expansion device 140 may be reduced. Accordingly, in the present embodiment, the heat exchanger connecting pipe 260 is disposed between the first and second heat exchangers 210 and 250 to solve this problem.

4 is a graph showing the relationship between the amount of consumed energy and the change in the compressor suction temperature according to the length of the first heat exchanger 210, that is, the length of the first suction heat exchanger 211 or the length of the condenser heat exchanger 213 Show the value.

As the length of the first heat exchanger 210 increases, that is, as the amount of heat exchange in the first heat exchanger 210 increases, the heat absorbed by the refrigerant sucked into the compressor 110 increases, The suction temperature is increased. Then, the amount of energy consumed by the operation of the deep-sea freezer 10 is reduced.

The suction temperature Ts of the compressor 110 may satisfy the following equation with respect to the ambient temperature (room temperature, To) based on the operating condition of the domestic compressor 110 according to the present embodiment.

To-5 ° C <Ts <To + 5 ° C

As the suction temperature Ts of the compressor 110 increases, the liquid refrigerant can be prevented from being sucked into the compressor 110, and the ineffective cooling power can be reduced. However, if the suction temperature Ts of the compressor 110 becomes too high, the discharge temperature or the discharge pressure of the compressor 110 may become too high.

As a result, in order to form an operating condition of the domestic compressor 110 according to the present embodiment and an appropriate level of compressor discharge temperature, it is preferable that the suction temperature Ts of the compressor 110 satisfies the above equation .

Referring to FIG. 4, the length of the first heat exchanger 210 to satisfy the suction temperature (Ts) condition of the compressor 110 may be about 3.5 to 5 m. That is, if the length condition of the first heat exchanger 210 is satisfied, the operating condition of the domestic compressor 110 according to the present embodiment can be satisfied and the operation reliability of the compressor can be improved.

The pipe diameter of the condensing heat exchanger 213 may be larger than the pipe diameter of the expansion device 140. For example, the pipe diameter of the condensation heat exchanger 213 may be in the range of 3.5 to 4.5 times the pipe diameter of the expansion device 140. In detail, the pipe diameter of the condensing heat exchanger 213 is 3.5 mm, and the pipe diameter of the expansion device 140 is 0.8 mm.

The refrigerant passing through the condensing heat exchanger 213 must be condensed. On the other hand, the refrigerant passing through the expansion device 140 must be decompressed. Actually, referring to the P-h diagram of FIG. 3, the dryness of the outlet state (point 4) of the expansion device 140 is formed higher than the dryness of the inlet state (point 3). That is, in the course of the refrigerant passing through the expansion device 140, vaporization occurs together with the decompression.

In summary, in the expansion device 140, since the refrigerant is intended to be decompressed, the diameter of the pipe can be reduced to increase the flow rate of the refrigerant, thereby reducing the pressure of the refrigerant.

On the other hand, if the pipe diameter of the condensing heat exchanger 213 is too small, the condensing heat exchanger 213 acts as a resistance against the refrigerant, thereby reducing the refrigerant flow rate and reducing the refrigerant pressure. On the other hand, Lt; / RTI &gt;

Therefore, in this embodiment, the pipe diameter of the condensing heat exchanger 213 is formed to be sufficiently larger than the pipe diameter of the expansion device 140 so that the condensation heat exchanger 213 does not act as a resistance against the refrigerant . In this case, the relatively large vapor-phase refrigerant can easily flow through the condensing heat exchanger 213 and can be sufficiently condensed in the process of heat exchange in the first heat exchanger 210.

Hereinafter, the second to fifth embodiments of the present invention will be described. These embodiments differ from the first embodiment only in some configurations, and therefore, differences will be mainly described. The same parts as those of the first embodiment are described with reference to the first embodiment and the reference numerals.

6 is a view illustrating a refrigeration cycle of the deep-sea freezer according to the second embodiment of the present invention.

Referring to Figure 6, the deep temperature freezer 10a according to the second embodiment of the present invention includes a compressor 110, a condenser 120, a dryer 130, an expansion device 140, And a condensation pipe 161 extending to the expansion device 140 is included.

The deep heat freezer 10a includes a first heat exchanger 210 for performing heat exchange between the condensed refrigerant and the suction refrigerant to the compressor 110 and a second heat exchanger 210 for performing heat exchange between the refrigerant passing through the expansion device 140 and the suction refrigerant And a second heat exchanger (250) for performing the second heat exchange.

The description of the above configurations is based on the description of the first embodiment.

The deep-sea freezer (10a) further includes a plurality of evaporators (151, 152) for evaporating the refrigerant decompressed in the expansion device (140).

The plurality of evaporators 151 and 152 include a first evaporator 151 installed on an outlet side of the expansion device 140 and a second evaporator 152 installed on an outlet side of the first evaporator 151 . The first and second evaporators 151 and 152 may be connected in series.

The deep-sea freezer (10a) may include a plurality of storages corresponding to the plurality of evaporators (151, 152). The plurality of storage chambers may include a cryogenic storage chamber at -60 ° C or lower and a freezer chamber at about -20 ° C. For example, the cool air generated by the first evaporator 151 may be supplied to the cryogenic storage chamber, and the cool air generated by the second evaporator 152 may be supplied to the freezer chamber.

The second heat exchanger 250 may be installed at the outlet of the second evaporator 152 and the first heat exchanger 210 may be installed at the outlet of the second heat exchanger 250. The refrigerant evaporated in the second evaporator 152 is absorbed while passing through the second heat exchanger 250 and the first heat exchanger 210. As a result, the temperature of the refrigerant sucked into the compressor 110 increases, Can be achieved.

7 is a view showing a refrigeration cycle of the deep-sea freezer according to the third embodiment of the present invention.

Referring to Figure 7, the deep temperature freezer 10b according to the third embodiment of the present invention includes a compressor 110, a condenser 120, a dryer 130, an expansion device 140, And a condensation pipe 161 extending to the expansion device 140 is included.

The expansion device 140 includes two expansion devices. The two expansion devices are provided with a first expansion device 141 capable of flowing at least a portion of the refrigerant flowing through the condensing pipe 161 and a second expansion device 141 connected in parallel with the first expansion device 141, And a second expansion device (143) through which another portion of the refrigerant flowing through the first expansion device (161) can flow.

The condensing pipe 161 is provided with a valve device for introducing the refrigerant flowing through the condensing pipe 161 into the expansion device 143 of at least one of the first expansion device 141 and the second expansion device 143, (170) may be installed. For example, the valve apparatus 170 may include a three-way valve. The condensing pipe 161 is connected to the inlet of the three-way valve, and the first and second expansion devices 141 and 143 may be respectively connected to the two outlet portions of the three-way valve.

The deep heat freezer 10b is provided with a first evaporator 151a connected to the outlet side of the first expansion device 141 and a second evaporator 152a connected to the outlet side of the second expansion device 143 .

The deep heat freezer 10b is further provided with a first evaporation pipe 181 extending from the first outlet of the valve device 170 to the first evaporator 151a and a second evaporation pipe 182 extending from the second outlet of the valve device 170 to the first evaporator 151a, And a second evaporation pipe 183 extending from the second evaporator 152a to the second evaporator 152a.

The first evaporation pipe 181 and the second evaporation pipe 183 may be joined together at the joint part 185. The joint portion 185 may be a point of the first evaporation pipe 181 or the second evaporation pipe 183. [ According to this configuration, the first evaporator 151a and the second evaporator 152a may be connected in parallel.

The first evaporation pipe 181 may be provided with a check valve 158 for guiding the unidirectional flow of the refrigerant in the first evaporation pipe 181. The refrigerant flow from the joint portion 185 toward the first evaporator 151a can be restricted by the check valve 158. [ As a result, the refrigerant having passed through the second evaporator 152a can be prevented from flowing into the first evaporator 151a through the joint portion 185. [

At least one evaporator of the first and second evaporators 151a and 152a may be operated under the control of the valve device 170. Only the refrigerant flow from the valve device 170 to the first evaporator 151a may be generated when the first outlet of the two outlet portions of the valve device 170 is opened and the second outlet portion is closed.

On the other hand, when the second outlet of the two outlet portions of the valve device 170 is opened and the first outlet portion is closed, only the refrigerant flow from the valve device 170 to the second evaporator 151a may be generated .

Of course, when both outlet portions of the valve device 170 are opened, the refrigerant introduced into the valve device 170 branches to the first and second evaporators 151a and 152a through the first and second outlet portions, .

The deep-sea freezer (10b) may include a plurality of storages corresponding to the plurality of evaporators (151a, 152a). In the plurality of storage chambers, two cryogenic storage chambers below -60 캜 may be included. As another example, the plurality of storage chambers may include a cryogenic storage chamber at -60 캜 or lower and a freezer chamber at about -20 캜.

The refrigerant having passed through the first evaporator 151a or the second evaporator 152a may pass through the second heat exchanger 250a. The second heat exchanger 250a is provided with at least a part of the first expansion device 141, the second expansion device 143 and the suction pipe 165, that is, the second suction heat exchanger 251 described in the first embodiment ) May be included.

The first and second expansion devices 141 and 143 and the second suction heat exchanging part 251 may be arranged to be in contact with each other. For example, the first and second expansion devices 141 and 143 and the second suction heat exchanging part 251 may be coupled by soldering.

A first heat exchanger 210a may be installed at the outlet side of the second heat exchanger 250a. The first heat exchanger 210a is provided with at least a portion of the condensing pipe 161, that is, at least this portion of the condensing heat exchanger 213 and the suction pipe 165 described in the first embodiment, A heat exchange unit 211 may be included. The description of the actions of the first heat exchanger 210a and the second heat exchanger 250a is based on the description of the first embodiment.

FIG. 8 is a view showing a refrigeration cycle of the deep-sea freezer according to the fourth embodiment of the present invention.

8, the deep temperature freezer 10c according to the fourth embodiment of the present invention includes a compressor 110, a condenser 120, a dryer 130, an expansion device 140, And a condensation pipe 161 extending to the expansion device 140 is included.

The deep heat freezer 10c includes a first heat exchanger 210 for performing heat exchange between the condensed refrigerant and the suction refrigerant to the compressor 110 and a second heat exchanger 210 for performing heat exchange between the refrigerant passing through the expansion device 140 and the suction refrigerant And a second heat exchanger (250) for performing the second heat exchange.

The description of the above configurations is based on the description of the first embodiment.

The deep-sea freezer (10c) further includes a plurality of evaporators (151b, 152b, 153b) for evaporating the refrigerant decompressed in the expansion device (140).

The plurality of evaporators 151b, 152b and 153b are provided with a first evaporator 151b installed on the outlet side of the expansion device 140 and a second evaporator 151b installed on the outlet side of the first evaporator 151b And a third evaporator 153b installed on the outlet side of the second evaporator 152b. The first, second and third evaporators 151b, 152b and 153b may be connected in series.

The deep-sea freezer 10c may include a plurality of storages corresponding to the plurality of evaporators 151b, 152b, and 153b. The plurality of storage chambers may include a cryogenic storage chamber at -60 ° C or lower, a freezing chamber at about -20 ° C, and a refrigerating chamber at 0 to 5 ° C. For example, the cool air generated by the first evaporator 151b may be supplied to the cryogenic storage chamber, the cool air generated by the second evaporator 152b may be supplied to the freezer chamber, and the third evaporator 153b May be supplied to the refrigerator compartment.

The second heat exchanger 250 may be installed on the outlet side of the third evaporator 153b and the first heat exchanger 210 may be installed on the outlet side of the second heat exchanger 250. The refrigerant evaporated in the second evaporator 152 is absorbed while passing through the second heat exchanger 250 and the first heat exchanger 210. As a result, the temperature of the refrigerant sucked into the compressor 110 increases, Can be achieved. The description related to this is based on the description of the first embodiment.

9 is a view showing a refrigeration cycle of the deep-sea freezer according to the fifth embodiment of the present invention.

Referring to FIG. 9, the deep-sea freezer 10d according to the fifth embodiment of the present invention includes two independent refrigeration cycles. The configurations of the two independent refrigeration cycles are the same.

The two refrigeration cycles include a first refrigeration cycle. In detail, the first refrigeration cycle includes a first compressor 110a, a first condenser 120a, a first dryer 130a, a first expansion device 140a, a first condensation pipe 161a, a first evaporator 150a, a first suction pipe 165a, a second heat exchanger 250b, and a first heat exchanger 210b. The description of these configurations and actions is the same as that of the first embodiment.

The two refrigeration cycles include a second refrigeration cycle. In detail, the first refrigeration cycle includes a second compressor 110b, a second condenser 120b, a second dryer 130b, a second expansion device 140b, a second condensation pipe 161b, a second evaporator 150b, a second suction pipe 165b, a fourth heat exchanger 250c, and a third heat exchanger 210c. The description of these configurations and actions is the same as that of the first embodiment.

According to the present embodiment, two independent refrigeration cycles are operated to cool a plurality of storage rooms provided in the deep-sea freezer 10d. In the plurality of storage chambers, two cryogenic storage chambers below -60 캜 may be included. The cool air generated in the first refrigeration cycle may cool the first cryogenic storage chamber, and the cool air generated in the second refrigeration cycle may cool the second cryogenic storage chamber.

110: compressor 120: condenser
130: dryer 140: expansion device
150: Evaporator 161: Condensation piping
165: suction pipe 170: valve device
181: first evaporation pipe 183: second evaporation pipe
185: joint part 210: first heat exchanger
250: second heat exchanger

Claims (17)

A compressor for compressing two or more mixed refrigerants;
A condenser for condensing the mixed refrigerant compressed in the compressor;
An expansion device for decompressing the mixed refrigerant condensed in the condenser;
An evaporator for evaporating the mixed refrigerant decompressed in the expansion device;
A condensing pipe extending from the outlet side of the condenser to the expansion device and guiding the flow of the mixed refrigerant;
A suction pipe extending from the outlet side of the evaporator to the compressor and guiding the suction of the mixed refrigerant into the compressor; And
And a plurality of heat exchangers installed in the suction pipe and performing heat exchange of the mixed refrigerant sucked into the compressor. The method according to claim 1,
Wherein the plurality of heat exchangers includes a first heat exchanger,
In the first heat exchanger,
A first suction heat exchanger for guiding the flow of the mixed refrigerant sucked into the compressor; And
And a condensing heat exchanging part for performing heat exchange with the first suction heat exchanging part and guiding the flow of the condensing pipe. 3. The method of claim 2,
Wherein the length of the first suction heat exchanger or the condensation heat exchanger is in the range of 3.5 m to 5 m. 3. The method of claim 2,
The pipe diameter of the condensation heat-
Wherein a diameter of the pipe of the expansion device is larger than a diameter of the pipe of the expansion device. 5. The method of claim 4,
The pipe diameter of the condensation heat-
Is formed within a range of 3.5 to 4.5 times the pipe diameter of the expansion device. 3. The method of claim 2,
The plurality of heat exchangers include a second heat exchanger,
In the second heat exchanger,
A second suction heat exchanger provided at one side of the first suction heat exchanger and guiding the flow of mixed refrigerant sucked into the compressor; And
And the expansion device performing heat exchange with the second suction heat exchanger. The method according to claim 6,
The first suction heat exchanger and the condensing pipe, or
The second suction heat exchanger, and the expansion device,
Wherein the heat exchanger is brought into contact with each other to perform heat exchange. 8. The method of claim 7,
The first suction heat exchanger and the condensing pipe, or
The second suction heat exchanger, and the expansion device,
Wherein the heat sink is coupled by soldering. The method according to claim 6,
Further comprising a heat exchanger connecting pipe disposed between the first and second heat exchangers for preventing heat exchange between the expansion device and the condensation heat exchanger,
Wherein the first heat exchanging unit and the second heat exchanging unit are spaced apart from each other by the heat exchanger connecting pipe. The method according to claim 6,
Wherein the first heat exchanger is installed at an outlet side of the second heat exchanger based on a flow direction of the refrigerant flowing through the suction pipe. The method according to claim 6,
Wherein the second heat exchanger is installed on the outlet side of the first heat exchanger based on a flow direction of the refrigerant flowing through the condensing pipe. The method according to claim 1,
The evaporator includes a first evaporator and a second evaporator connected in series with each other,
Wherein the second evaporator is installed at an outlet side of the first evaporator. The method according to claim 1,
The evaporator includes a first evaporator and a second evaporator connected in parallel with each other,
Wherein the expansion device includes a first expansion device installed at an inlet side of the first evaporator and a second expansion device installed at an inlet side of the second evaporator. 14. The method of claim 13,
Wherein the first and second expansion devices and the suction pipe are coupled to each other to perform heat exchange. The method according to claim 1,
In the evaporator,
A first evaporator installed at an outlet side of the expansion device;
A second evaporator connected in series to an outlet side of the first evaporator; And
And a third evaporator connected in series to an outlet side of the second evaporator. The method according to claim 1,
Two independent refrigeration cycles are driven,
Wherein each of the independent refrigeration cycles includes the compressor, the condenser, the expansion device, the evaporator, and a plurality of heat exchangers. The method according to claim 1,
Wherein the mixed refrigerant includes butane and ethylene.

Images