JPH03168566A - Operation of refrigeration cycle device - Google Patents

Abstract

PURPOSE:To enable a cycle operation to be attained with a high result of performance using all kinds of refrigerants including mixed refrigerant only with a quite small amount of thermal physical value information by a method wherein a refrigeration cycle filling the refrigerant is operated in response to a specified condensing temperature. CONSTITUTION:New refrigerant, for example, a refrigerant replacing with the conventional refrigerant is filled in a refrigeration cycle. This refrigeration cycle is designed such that as to the refrigerant of which condensing temperature set in response to a limited temperature or boiling point of these refrigerants, i.e., the limited temperature can be made apparent, the value of about 90% of the limited temperature of the refrigerant is applied and as to the refrigerant having no limited temperature, a condensing temperature set in response to the value of about 90% of the temperature calculated by dividing a predetermined constant to the boiling point of the refrigerant. Under this operating condition, as the compressor 1 of the refrigeration cycle device is operated, theoretically, a refrigeration cycle operation having the highest coefficient of performance can be carried out.

Description

DETAILED DESCRIPTION OF THE INVENTION [Object of the Invention] (Industrial Application Field) The present invention relates to a method of operating a refrigeration cycle device such as a refrigerator or a heat pump using a vapor compression refrigeration cycle.

(Conventional technology) In the field of refrigeration and air conditioning engineering, which is typified by refrigerators and air conditioners, various research and developments have been carried out with the aim of improving storage capacity and comfort in living spaces. development can be seen.

In recent years, there has been an increase in interest in heat bombs, which pump heat from a low-temperature heat source to a high-temperature heat source, can use energy effectively, and are highly comfortable, and research and development related to heat bombs has increased. is being actively carried out. In particular, improvements in the performance and practicality of heat pumps using vapor compression refrigeration cycle devices over the past decade have been extremely remarkable, and heat pumps have come to be widely used in home and consumer air conditioning systems. Furthermore, for systems that require relatively high temperatures, methods that directly utilize combustion heat or electric heaters have traditionally been used, but as the performance of heat pumps has improved, the upper limit of the usable temperature that can be obtained with heat pumps has increased. Since it has reached approximately 430K and the cycle efficiency (coefficient of performance) of heat pumps has become sufficiently higher than conventional boiler efficiency and heater efficiency,
Heat pumps are increasingly being used. In addition, heat pumps are suitable for recovering and reusing waste heat, which is a relatively low-temperature heat source.
Heat bombs are also used in various waste heat recovery systems.

However, the performance of such a refrigeration cycle device, like conventional systems, depends on the performance of the component equipment and working fluid.

Therefore, one way to improve the performance of refrigeration cycle equipment is to optimize the system. Of course,
In this case, information on the thermophysical properties of the working fluid regarding the refrigeration cycle becomes essential information.

Substances used or considered usable in the working fluid of refrigeration cycle equipment include chlorofluorocarbon (chlorofluorocarbon) refrigerants, hydrocarbon refrigerants, fluorinated alcohols, and inorganic compounds. Currently, fluorocarbon-based refrigerants are widely used from the comprehensive viewpoint of safety, price, chemical stability, and impact on cycle component equipment materials.Especially used in heat pumps. There are more than 20 types of fluorocarbon refrigerants that can be used, but among them, R
1. Two types, 2 and R22, have been widely used as refrigerants (working fluids).

However, in recent years, the range of application of heat pumps has expanded and it has become necessary to apply them to various forms of application, and the critical temperatures of R12 and R22 have increased to 385.01 K and 369 K.
．． 32"K, and from their thermodynamic properties near their critical points, R12 and R22 are t of the high temperature heat bomb.
It is not suitable as a working fluid. Therefore, R1
We are currently considering a high-boiling refrigerant with a relatively high critical point, such as R14, or a two-component refrigerant, which is a combination of a high-boiling refrigerant and a refrigerant with a relatively low boiling point, such as R12 or R22.

This two-component mixed refrigerant can be treated in the same way as a single refrigerant, and there are two types: azeotropic refrigerant mixtures with an azeotropic point, and non-azeotropic refrigerant mixtures whose thermodynamic properties change depending on the component composition and do not have an azeotropic point. being classified. In particular, the effectiveness of using a non-azeotropic refrigerant mixture as the working fluid of a heat pump has been pointed out. In other words, when considering the heat transfer properties between the high-temperature heat source, the low-temperature heat source, and the heat exchanger between the cycles, the Lorenz cycle, which is said to have a higher cycle efficiency than the reverse Carnot cycle, can be compared to a non-azeotropic cycle. This can be achieved using a mixed refrigerant. Also, when the heat transfer characteristics in the heat exchanger are not taken into account, there is no difference! The cycle characteristics of the 1i mixed refrigerant, which can be discussed only based on the thermodynamic properties of the working fluid, show that the coefficient of performance improves compared to cycles that use the component materials as the working fluid, and that the cycle characteristics improve when the components are changed. It has been pointed out that non-azeotropic refrigerants are effective in improving the diversity of properties.

However, in recent years, there has been a fear that refrigerants with these characteristics, and even refrigerants that are suitable for operating conditions, may no longer be available due to fluorocarbon regulations.

In other words, in recent years, due to the chemical stability that has traditionally been a characteristic of fluorocarbon refrigerants, the depletion of the ozone layer in the atmosphere and the effect on the greenhouse effect due to fluorocarbon refrigerants released into the atmosphere have become a problem. There is.

? For this reason, the issue of ozone layer depletion due to fluorocarbon refrigerants used in aerosols and refrigerants released into the atmosphere, the so-called fluorocarbon/ozone problem, was raised as a springboard, and the impact of fluorocarbons on the global environment has been raised. Chemical evaluations began to be reviewed, and fluorocarbon regulations began to be implemented.

Under fluorocarbon regulations, R12 does not have hydrogen in its molecular structure, so it is classified as a chlorofluorocarbon (CFC)-based refrigerant and is subject to regulations. Furthermore, even R22, which does not belong to CFC-based refrigerants, is said to destroy the ozone layer, although the amount of destruction is smaller than that of CFC-based refrigerants, and it is currently mainly used as a working fluid in refrigeration cycle equipment (heat pumps, refrigeration systems, etc.). Most of the fluorocarbon-based refrigerants currently in use are subject to fluorocarbon regulations and may become unusable by the end of this century.

However, fluorocarbon refrigerants play a major role in the current field of refrigeration and air conditioning engineering.

In response to this, there is an increasing demand for the development of alternative refrigerants and non-azeotropic merging refrigerants, and for systematic elucidation of their thermophysical properties and other properties.

(Problem to be solved by the invention) However, at present, information on the thermophysical properties of not only alternative refrigerants but also binary non-azeotropic refrigerants with a short history is very limited compared to information on conventional single refrigerants. limited to. For example, when designing a cycle using a binary azeotropic refrigerant mixture, it is necessary to select a working fluid suitable for the cycle, set the mixing ratio, cycle conditions, etc., and there is a wide range of working fluids to be used. Although it is necessary to obtain highly accurate thermophysical property value information in a temperature/pressure range, such information is not available, and conventionally most cases have been based on thermophysical property value information based on simple estimation formulas.

For this reason, even if new refrigerants, alternative refrigerants, or two-component mixed refrigerants are developed, cycle design must be performed based on insufficient information on thermophysical properties, and it is difficult to operate a refrigeration cycle that achieves a very high coefficient of performance. It was inadequate. In particular, there is little information on the thermophysical properties of new refrigerants, alternative refrigerants, and binary mixed refrigerants, and it is impossible to obtain optimal cycles using these refrigerants. It is necessary to conduct actual machine tests that require

This invention was made with attention to these circumstances,
The purpose of this study is to develop a method for operating refrigeration cycle equipment that can achieve cycle operation with a high coefficient of performance using all kinds of refrigerants, including mixed refrigerants, with only a small amount of information on thermophysical properties. Our goal is to provide the following.

[Yokohama of the Invention] (Means for Solving the Problems and Their Effects) In order to achieve the above object, the method for operating a refrigeration cycle device of the present invention is such that a refrigerant whose critical temperature is known is an abbreviation of the critical temperature of the refrigerant. For a refrigerant whose critical temperature is unknown and whose boiling point is known, set the condensing temperature at a value approximately 90% of the temperature determined by dividing the boiling point temperature of the refrigerant by a predetermined constant, and set the condensing temperature according to this condensing temperature. A refrigeration cycle filled with the refrigerant is operated.

As a result of repeated trial and error studies, if the critical temperature or boiling point temperature of the refrigerant is clear, the above process can be applied to it to set the condensing temperature value, and the refrigeration cycle can be operated based on this condensing temperature. Easy to use without requiring any information other than thermophysical property information! This is the one that can obtain the highest coefficient of performance for n.

(Example) The present invention will be described below based on a first example shown in FIGS. 1 to 10.

FIG. 1 shows a vapor compression type refrigeration cycle device (reverse Rankine cycle) to which the present invention is applied, where 1 is a compressor driven by, for example, a motor la (drive section), 2 is a condenser, and 3 is a compressor driven by a motor la (drive section).
is an expansion device composed of, for example, an expansion valve; 4 is an evaporator;
6 is the preceptor.

These devices are connected in order through a refrigerant path 5 to form a refrigeration cycle.

This refrigeration cycle is filled with, for example, a new refrigerant or an alternative refrigerant to replace the conventional refrigerant (both are considered to be a single refrigerant). In addition, this refrigeration cycle has a condensing temperature set based on the critical temperature or boiling point of these refrigerants. The boiling point temperature (absolute temperature) of this refrigerant is r
O. It is designed to operate according to a condensing temperature set at a value of 90% of the temperature found by dividing 7 (predetermined constant).

However, if the compressor 1 of the refrigeration cycle device is operated under these operating conditions, as a result of repeated trial and error studies, the refrigeration cycle operation with the highest coefficient of performance can theoretically be achieved for the following reasons. I found out that I can get it.

The results of this study will be described below.The study began by considering a refrigeration cycle using a conventional refrigerant.

First, we started by examining a theoretical cycle that can be discussed depending only on the thermodynamic properties of the working fluid.

The refrigeration cycle studied here was a basic vapor compression reverse Rankine cycle having the structure shown in FIG. In this refrigeration cycle, the heat transfer coefficient of the heat exchanger (condenser 2, evaporator 4) is infinite, the adiabatic efficiency of the compressor 1 is 100%, and there is no cycle characteristic due to pressure loss. It is a theoretical cycle.

This is because when creating an actual refrigeration cycle device, the system is constructed with the aim of obtaining a coefficient of performance close to the coefficient of performance of this theoretical cycle. In other words, the coefficient of performance of a theoretical cycle is extremely important for optimal cycle design.

Fig. 2 shows the rT-s diagram of this refrigeration cycle, and Fig. 3 shows the same p-h diagram.

Here, when the compression work in the compressor 1 in this theoretical cycle is rWJ and the heat radiation amount in the condenser 2 is rQJ, each compression work ffiW and heat radiation melon Q can be calculated as shown in the following equation by changing the symbol in Fig. 1. is required.

W-(hc-hB)-(1)Q=
(hc - hp) - (2) If we calculate the coefficient of performance cop of the heat bomb and the coefficient of performance COPc of the refrigerator from these, we can obtain the coefficient of performance COP of the heat pump and the coefficient of performance CPO of the refrigerator in the refrigeration cycle equipment.
c becomes as follows.

C O P = Q / W − 1 + C P O
c...(3) Next, let us consider conventionally used refrigerants.

Here, we will focus on a single refrigerant.

Conventionally, R12 and R2 are mainly used as refrigerants.
2, R114, which is attracting attention as a high boiling point refrigerant,
R, which contains bromine as a component element and is said to be suitable for low temperatures.
13B1, was applied to R502, which is used as the low-temperature side working fluid of binary refrigerators. In addition to R15'2a, which is attracting attention as an alternative refrigerant, and fluorocarbon-based refrigerants, we also focused on ammonia, which was used in early refrigerators, and propane, which was used in cold power generation and industrial complexes.

Then, the performance of a heat pump cycle using these refrigerants was calculated based on the above equation of the theoretical cycle, and when looking at the calculation results, a diagram of the coefficient of performance as shown in FIG. 4 appeared.

Furthermore, from the above equation (3), the behavior of the coefficient of performance of heat pumps and refrigerators is basically the same, just different levels, so for convenience of explanation, heat pumps are taken as an example. Let's take Heat Bomb as an example.

In the above figure 4, R12, R22, R13B1, R50
2. The relationship between each coefficient of performance COP and condensation temperature T con of a single fluorocarbon refrigerant of R114 is shown. Specifically, FIG. 4 shows the supercooling degree Tsc at the inlet of the expansion device 3.
(shown in FIG. 2) is "5" K, the degree of superheating Tsh (shown in FIG. 2) of the compressor 1 is "0" K, and the condensation temperature of the heat bomb cycle (T co
Calculate by setting the temperature increase width, which is the temperature difference between n) and the evaporation temperature (T eva), to a constant value, for example, "30" K, "60" K, and "90 #K, and calculate each temperature increase width at this time. and isobar lines corresponding to the boiling point pressure are shown for each of the above five types of refrigerants.

Looking at this FIG. 4, no regularity is seen in the coefficient of performance COP of the heat pump cycle using each refrigerant.

However, it has been found that it is practical to use the reduced condensation temperature (T con) r, which is the condensation temperature Tsh divided by the critical temperature (absolute temperature). The results are shown in FIG.

That is, looking at Figure 5, when the temperature increase range of each refrigerant is constant, the maximum coefficient of performance COP for each refrigerant is that the condensation temperature is the converted condensation temperature (T con) r is about 0.9.
It is obtained by As a result, the coefficient of performance COP of each refrigerant
We found regularities about As shown in FIG. 6, this regularity was the same for R152a, R22, ammonia, and propane gas.

This means that as long as you know the critical temperature of the refrigerant as thermophysical property value information, regardless of what kind of refrigerant it is,
If you set the condensation temperature to 90% of the critical temperature of the refrigerant, determine the operating conditions, and operate the refrigeration cycle to achieve this condensation temperature, you will achieve the highest precept! The j2 coefficient becomes COP.

Furthermore, looking at FIGS. 5 and 6, although the behavior of ammonia is slightly different due to the difference in molecular composition, the boiling point pressure of each refrigerant is found to be regular with "about 0.7". From the boiling point pressure in the same figure, the boiling point is rO. 7X critical temperature".

This is also proven from the following principle.

That is, there is a principle called the two-variable correspondence state principle in statistical mechanics. For molecules that fully satisfy this condition, the following formula is applied.

Fog(P/PC)Tr-0.7”t.oo
However, P is boiling point pressure and Pc is critical pressure.

According to this equation, r P / P c = 0.
When I J, Tr=0.7J. However, 2
The variable-corresponding state principle is limited to simple molecules such as Ay + K'* CH4, but when applied to more complex molecules, the effects of the size and polarity of the molecule cannot be ignored. A deviation of 30% is observed.

Therefore, we rearranged using the parameter called eccentricity coefficient ω,
Using the three-variable correspondence principle, we get the following.

ω− j2og(P/PC)r,. o. 71.
00 Generally, the value of the eccentricity coefficient is 0 to 0.3 J, and in the case of molecules with strong polarity, the value is higher.

Therefore, when the molecular structure is complex, rP/Pc≦0,
1, it becomes TrxQ, 7J.

For example, the Pc values of R12, R22, and R114 are r4.129, r4.990J, and 3.252, respectively.
JMPa, rTr-0.7J is Pc, respectively.
The boiling point of ``rl/10 or less'' (this is how the converted boiling point temperature for most refrigerants is ``around 0.7'', and the regularity of ``about 0.7'' can be found from this.

With this constant expressed by rTr-0.7J, for example, for a refrigerant whose critical state quantity is not clarified, that is, a refrigerant whose critical temperature is unknown, the highest coefficient of performance when the temperature rise width is constant is ．． From the regularity of 9J, it can be seen that the boiling point temperature can be obtained at a converted boiling point temperature of about 0.9/0.7J.Of course, in Figures 5 and 6, the curves with a constant temperature increase range are the same for each refrigerant. Since they have the same shape, if you know the coefficient of performance under a certain cycle setting condition, you can also know the coefficient of performance under other conditions based on the critical temperature or boiling point temperature.

On the other hand, when looking at FIGS. 7 and 8 with attention to the temperature increase range and the degree of heating of the compressor 1, the following was found.

That is, Fig. 7 shows the relationship between the coefficient of R12 and the condensing temperature Tcon, and Fig. 8 shows the case where the heating degree of the compressor 1 is set to a large value of "20" K. .

In FIG. 7, the dashed-dotted line represents the highest coefficient of performance when the temperature increase width is constant. Paying attention to this dashed-dotted line, as the temperature increase width increases, the performance coefficient COP when the temperature increase width is constant becomes the condensation R degree, rO. It is obtained at a reduced condensation temperature of 85J. In addition, as shown in Figure 8, the heating degree of the compressor population is increased to "20K".
Then, the coefficient of performance when the heating radiation is constant is the seventh
In the case shown in the figure as well, the condensation temperature can be obtained at a high condensation temperature of "10" K.

However, as a result of many such studies, we found that under most cycle conditions using various refrigerants, the highest coefficient of performance COP when the temperature increase width is constant is the equivalent condensing temperature rO. It has been confirmed that it can be obtained at 85 to 0.94 J, and it can be said that a high coefficient of performance COP can be maintained if the converted condensation temperature as described above is "about 0.9".

Thus, even if the information on the thermophysical properties of a refrigerant is limited, if the critical temperature or boiling point temperature is clarified as the minimum information (or if it is clarified), the value of approximately 90% of the refrigerant's critical temperature can be determined. ” or “0 to the boiling point temperature of the refrigerant.”
．． It can be seen that the condensing temperature at which the highest coefficient of performance COP of the refrigerant can be obtained can be approximately predicted by using the value that is approximately 90% of the temperature obtained by subtracting 7.

And, this can be applied as it is to most other types of refrigerants, such as the new refrigerants and alternative refrigerants mentioned above.

Therefore, by operating the refrigeration cycle of a heat pump (or refrigerator) according to the operating conditions, it is possible to easily achieve the highest coefficient of performance COP, regardless of whether a new refrigerant or an alternative refrigerant is used. .

In addition to the above, it was also confirmed from FIGS. 9 and 10 that if the population specific volume of the compressor is known in addition to the critical temperature or boiling point, an optimal cycle design is possible.

In other words, Fig. 9 shows the relationship between (Q/Vl) and condensing temperature for heat pump cycles using various single refrigerants, as the condensing pressure rises to ``30''K1 ``60''K1 ``90''K. The corresponding condensation temperatures are shown. In addition, heat radiation ff
The ratio (Q/Vt) between iQ and compressor inlet specific volume v1 is important in cycle design such as determining the volume of the compressor 1 and the heat exchanger.

Here, in this FIG. 9, no regularity is seen in the values of (Q/vl) between the respective refrigerants. However, when the condensation temperature is divided by the critical temperature for conversion, regularity appears as shown in FIG. This regularity holds true for most refrigerants, with the exception of substances with significantly different molecular structures such as ammonia. That is, the heat radiation amount in the condenser 2 is organized by the compressor inlet specific volume v1 and the critical temperature. In other words, if the compressor inlet specific volume v1 and critical temperature or boiling point temperature are known, the amount of heat released by many refrigerants can be roughly estimated, and an optimal cycle design can be made based on this.

In the first embodiment, the present invention is applied to a refrigeration cycle of a heat pump (or refrigerator) using a single refrigerant, but it can of course also be applied to a refrigeration cycle using a mixed refrigerant.

Examples of cycles using this mixed refrigerant are shown in FIGS. 11 to 16 as a second embodiment.

Before explaining how to set the condensation temperature of a mixed refrigerant, it is necessary to explain the mixed refrigerant cycle to which the mixed refrigerant is applied, which is the premise.Since the mixed refrigerant cycle is different from the above single refrigerant conditions, the standard Assuming that the phase changes in the condenser 2 and evaporator 4 both occur under constant pressure, the cycle will be a "T-sJ diagram" as shown in FIGS. 11 and 12.

By the way, two cases can be assumed for setting the temperature conditions of this mixed refrigerant cycle, since the slopes of each change in the condenser 2 and the evaporator are approximately equal.

One is when the condenser inlet temperature and evaporator outlet temperature shown in Fig. 11 are determined (temperature conditions convenient for organizing the temperature increase width), and the other is when the condenser population temperature and evaporator temperature are determined. This is the case when the inlet temperature is determined (temperature conditions that provide the most ideal and high performance coefficient COP).

Cycle setting conditions will be selected from these.

Assuming an actual heat pump, if a heat exchanger with a sufficiently large heat transfer area is used, for example, a counter-current heat exchanger, the conditions corresponding to those shown in FIG. 12 are considered to be close, simply from the viewpoint of the structure of the heat exchanger. Specifically, for example, assuming a heat pump that uses air as a medium, the required hot air target temperature corresponds to the temperature of the air coming out of the condenser, and is set at a temperature higher than the temperature at which frost forms in the evaporator. This is because it corresponds to the temperature of the air coming out of the evaporator.

However, if the condenser is indoors and the evaporator is outdoors, the indoor set temperature is generally set to the average temperature of the condenser, and since the outside air temperature does not change much, the evaporator inlet temperature is set. .

Then, the cycle setting conditions are not as shown in Fig. 12 but as shown in Fig. 11.
Closer to the picture. In particular, if the hot air blown out from the condenser is set to the set temperature, the cycle setting conditions will be as shown in FIG. 11.

In addition, in the case of a refrigerator, for example, assuming a freezer, and assuming that the frozen rhinoceros uses a counter-flow heat exchanger, the evaporator is inside the refrigerator, and the condenser is outside the refrigerator, the set temperature inside the refrigerator is is generally set to the average temperature of the evaporator. Furthermore, since the temperature outside the refrigerator varies greatly, the condenser temperature is set, and the cycle setting conditions are close to those shown in FIG. 11, similar to the heat pump described above.

However, a heat bomb for air conditioning or a heat bomb for hot water supply, which has an infinite heat source as a low temperature side, has a high temperature side with a finite area or amount of heat, and determines the cycle characteristics by setting the temperature of the medium coming out of the condenser. In this case, or in the case of a refrigerator which is the opposite, the cycle setting conditions shown in FIG. 11 are suitable. Of course, this is limited to cycle analysis that only considers the thermodynamic properties of the working fluid.

As an example of the mixed refrigerant, we took up the "rR22+R114 non-azeotropic mixed refrigerant" and examined its cycle characteristics according to the cycle setting conditions shown in FIG. Understood.

That is, FIG. 13 shows rR12+ with different component compositions.
The relationship between the coefficient of performance COP and the condenser artificial temperature TD of five types of refrigerants of the R114 series is shown. Specifically, the first
In Figure 3, the supercooling degree Tsc is “5”K, and the superheating degree Tsh
is set to "0" K, and the temperature increase width of the heat bomb cycle is constant, for example, "30" K. Calculate by setting to "60"K and "90"K, and calculate each temperature rise width and R12 and R
The evaporator outlet temperatures corresponding to the boiling point pressures of 114 and 114 are respectively expressed.

Looking at this FIG. 13, no regularity is seen in the coefficient of performance COP of the heat pump cycle using each refrigerant.

However, when using the reduced condensation temperature (To)r obtained by dividing the condensation temperature T con by the critical temperature (absolute temperature), it was found that there is regularity as shown in FIG. 14.

In other words, as in the first embodiment, the highest coefficient of performance COP when the temperature increase width in each set is constant is obtained when the converted condensing temperature (To) r is "about 0.9" for each refrigerant. I know that it will happen. Also, from FIG. 14, as in the first embodiment, the boiling point is ro. 7X critical temperature," for example, the highest coefficient of performance COP when the temperature increase range of a refrigerant whose critical state quantity is not clarified is constant, is Similarly, “boiling point temperature X0.
It is thought that it can be obtained at about 9/0.7J. In the study, the same result was obtained from the coefficient of performance COP of the mixed refrigerant "rR12+R22 system" shown in Figure 10.

Here, the regularity based on the optimal coefficient of performance when the temperature increase range is constant is as same as in the case of a single refrigerant in the first embodiment, although substances with greatly different molecular structures such as ammonia differ. It can be said that this can be achieved with a mixed refrigerant.

Thus, even for mixed refrigerants, if the minimum information is to clarify the critical temperature or boiling point temperature (or
If it is clear), it can be seen that, as in the first embodiment, it is possible to roughly predict the condensing temperature at which the highest coefficient of performance COP of the mixed refrigerant is obtained.

However, even in the case of a mixed refrigerant, in which the cycle characteristics are changed to a desired degree by changing the composition to a certain extent (such as changing the target temperature), which type of refrigerant can be used will give the best result. You can see whether the performance coefficient can be obtained.

Thus, just by setting the component composition to the operating conditions, it is possible to easily obtain an optimal cycle that provides the highest coefficient of performance, as in the first embodiment.

In addition, the regularity regarding the ratio of the heat radiation amount Q to the compressor population specific volume V1 is also regular, as can be seen from the diagram of the converted condensing temperature shown in rR22+R114J shown in FIG. As in the example, if the compressor population specific volume, in addition to the critical temperature or boiling point, is known, an optimal cycle design is possible.

In the above-mentioned embodiment, the cycle setting in the case of a constant temperature increase width was described, but it is considered that the same effect can be brought about even in the case of I where the temperature increase width is not constant.

[Effects of the Invention] As explained above, according to the present invention, a high coefficient of performance can be achieved using all types of refrigerants, including new refrigerants, alternative refrigerants, and A-mix refrigerants, using only a small amount of thermophysical property value information. It is possible to realize a refrigeration cycle operation that brings about the following.

[Brief explanation of the drawing]

1 to 10 show a first embodiment of the present invention, FIG. 1 is a schematic configuration diagram showing a refrigeration cycle device using a single refrigerant to which this invention is applied, and FIG. 2 is a refrigeration cycle thereof. Figure 3 is the same <rP-hj diagram, Figure 4 is a line showing the relationship between the performance coefficient and condensation temperature in a heat pump refrigeration cycle using a fluorocarbon-based Qi refrigerant. Figure 5 is a diagram showing the relationship between the coefficient of performance and the converted condensing temperature in a heat pump refrigeration cycle using a single fluorocarbon refrigerant, and Figure 6 is a diagram showing the relationship between the coefficient of performance and the converted condensation temperature in a heat pump refrigeration cycle using a variety of different refrigerants. A diagram showing the relationship between the coefficient and the converted condensing temperature, Figure 7 is a diagram showing the relationship between the coefficient of performance and condensing temperature of a heat bomb refrigeration cycle using R12, and Figure 8 shows the same heating degree at "20". Figure 9 is a diagram showing the relationship between the coefficient of performance and the condensing temperature (Q) of a heat pump refrigeration cycle using a single fluorocarbon refrigerant
A diagram showing the relationship between /■1) and condensing temperature, Figure 10 shows the (
Figure 11 is an rT-sJ diagram of a refrigeration cycle using a mixed refrigerant according to the second embodiment of the present invention, and Figure 12 is a diagram showing the relationship between Q/V+) and converted condensing temperature. The condition "T-sJ diagram, Figure 13 is rR22+R1
Diagram showing the relationship between the coefficient of performance of a heat pump refrigeration cycle using a 14 series refrigerant mixture and the condenser population temperature, Figure 14 is rR22+R1 The coefficient of performance of a heat pump refrigeration cycle using a 14 series refrigerant mixture. A diagram showing the relationship between and converted condenser inlet temperature, Figure 15 is rR12 + R2
A diagram showing the relationship between the coefficient of performance of a refrigeration cycle of a heat pump using a 2-system mixed refrigerant and the converted condenser inlet temperature,
FIG. 16 is a diagram showing the relationship between (Q/V1) and the converted condensation degree d of a heat pump refrigeration cycle using a mixed refrigerant of "rR22+R1 14 system". 1... Compressor, 2... Condenser, 3... Expansion device,
4... Evaporator.

Claims (1)

[Claims] For refrigerants whose critical temperature is known, the value is approximately 90% of the critical temperature of the refrigerant, and for refrigerants whose critical temperature is unknown and whose boiling point temperature is known, the temperature is determined by dividing the boiling point temperature (absolute temperature) of the refrigerant by a predetermined constant. 1. A method of operating a refrigeration cycle apparatus, characterized in that a condensation temperature is set at a value of approximately 90%, and a refrigeration cycle filled with the refrigerant is operated according to this condensation temperature.