US20100024469A1

US20100024469A1 - Refrigerating cycle apparatus

Info

Publication number: US20100024469A1
Application number: US12/462,111
Authority: US
Inventors: Shun Kurata; Takashi Yamanaka; Shin Nishida; Yuuichi Shirota; Yasutaka Kuroda
Original assignee: Denso Corp
Current assignee: Denso Corp
Priority date: 2008-07-30
Filing date: 2009-07-29
Publication date: 2010-02-04
Also published as: JP2010032157A

Abstract

A refrigerating cycle apparatus includes a circulation refrigerant, a compressor, a condenser, an expansion valve, an evaporator, and an internal heat exchanger. The internal heat exchanger is disposed to perform heat exchange between a high pressure circulation refrigerant flowing from the condenser to the expansion valve and a low pressure circulation refrigerant flowing from the evaporator and the compressor. The circulation refrigerant has a property indicating an isentropic line having a gradient greater than a gradient of an isentropic line of a R134a refrigerant and a saturation characteristic curve having a two-phase region enthalpy width smaller than a two-phase region enthalpy width of the R134a refrigerant, on a p-h chart.

Description

CROSS REFERENCE TO RELATED APPLICATION

This application is based on Japanese Patent Application No. 2008-196316 filed on Jul. 30, 2008, the disclosure of which is incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to a refrigerating cycle apparatus

BACKGROUND OF THE INVENTION

Recently, new refrigerants focusing on global warming potential (GWP) have been proposed, as described in JP2008-505211A (U.S. Pat. No. 7,413,675), JP2008-111578A, JP2007-538115A (U.S. Pat. No. 7,524,805) and JP2007-535611A (U.S. Pat. No. 7,279,451). Although such new refrigerants have the advantage of low GWP, it is required to have a refrigerating capacity substantially equal to those of general refrigerants and to be easily substituted for the general refrigerants.
A refrigerating cycle apparatus described in JP2007-71461 (US2007/0074538) has an internal heat exchanger performing heat exchange between a high pressure refrigerant flowing between a condenser and an expansion valve and a low pressure refrigerant flowing between an evaporator and a compressor. The expansion valve has a normal charge characteristic that is capable of controlling a condition of refrigerant at an outlet of the evaporator such that the dryness thereof is 0.9 and the superheat degree is 5 degrees Celsius. Such a condition achieves a control biased to a liquid phase.

SUMMARY OF THE INVENTION

Although the refrigerating cycle apparatus described in JP2007-71461 has the internal heat exchanger, the temperature of refrigerant to be drawn into the compressor can be reduced. As such, even if the temperature of the refrigerant is increased inside the compressor, the temperature of the refrigerant discharged from the compressor is relatively low. Therefore, it is less likely that rubber parts, resin parts, electronic devices and the like on a periphery of the compressor will be affected by heat. Accordingly, deterioration of durability of such parts and devices is suppressed. On the other hand, if a gas and liquid two-phase refrigerant flows out from the evaporator, an opening degree of the expansion valve is likely to cause hunting. In refrigerating cycle apparatuses employing the general refrigerants, which has been widely used in markets, improvement of the refrigerating capacity has been attempted. However, there are new subjects arise with the improvement of the refrigerating capacity.
According to an aspect of the present invention, a refrigerating cycle apparatus includes a circulation refrigerant, a compressor, a condenser, an expansion valve, an evaporator, and an internal heat exchanger. The internal heat exchanger performs heat exchange between a high pressure circulation refrigerant flowing from the condenser to the expansion valve and a low pressure circulation refrigerant flowing from the evaporator to the compressor, the low pressure circulation refrigerant having pressure lower than that of the high pressure circulation refrigerant. The circulation refrigerant has a property indicating an isentropic line having a gradient greater than a gradient of an isentropic line of a R134a refrigerant and a saturation characteristic curve having a two-phase region enthalpy width smaller than a two-phase region enthalpy width of the R134a refrigerant, on a p-h chart, the two-phase region enthalpy width being defined between a saturated liquid line and a saturated gas line.
Since the refrigerating cycle apparatus has the internal heat exchanger, the circulation refrigerant can sufficiently exhibit its cooling capacity. Further, the temperature of the refrigerant can be reduced to a relatively low temperature region so that durability of component parts of the refrigerating cycle apparatus will not be deteriorated, while effectively exhibiting the property of the circulation refrigerant.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects, features and advantages of the present invention will become more apparent from the following detailed description made with reference to the accompanying drawings, in which like parts are designated by like reference numbers and in which:

FIG. 1 is a block diagram of a refrigerating cycle apparatus according to a first embodiment of the present invention;

FIG. 2 is a perspective view of the refrigerating cycle apparatus according to the first embodiment;

FIG. 3 is a block diagram of the refrigerating cycle apparatus for illustrating an expansion valve according to the first embodiment;

FIG. 4 is a chart showing a property of a refrigerant according to the first embodiment;

FIG. 5 is a chart showing a characteristic of the expansion valve according to the first embodiment;

FIG. 6 is a chart showing a relationship between a refrigerating capacity and a superheat degree according to the first embodiment;

FIG. 7 is a chart showing a refrigerating capacity and a discharge temperature according to the first embodiment;

FIG. 8 is a block diagram of a refrigerating cycle apparatus according to a second embodiment of the present invention;

FIG. 9 is a chart showing a characteristic of an expansion valve of the refrigerating cycle apparatus according to the second embodiment;

FIG. 10 is a block diagram of a refrigerating cycle apparatus according to a third embodiment of the present invention; and

FIG. 11 is a chart showing a characteristic of an expansion valve of the refrigerating cycle apparatus according to the third embodiment.

DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS

First Embodiment

A first embodiment will now be described with reference to FIGS. 1 to 7. A refrigerating cycle apparatus 100A of the present embodiment is exemplarily used for a vehicular air conditioner 100 for performing an air conditioning operation such as a cooling operation of a passenger compartment,.
Referring to FIGS. 1 and 2, an engine compartment 1 of a vehicle is separated from a passenger compartment 2 through a dash panel 3. The refrigerating cycle apparatus 100A is mounted in the engine compartment 1 and the passenger compartment 2 across the dash panel 3. An interior unit 100B of the air conditioner 100 is mounted in a space provided by an instrument panel in the passenger compartment 2.
The interior unit 100B has an air conditioning case 101. A blower 102, an evaporator 141, a heater core 103 and the like are disposed in the air conditioning case 101. The blower 102 selectively draws outside air and inside air and blows the air toward the evaporator 141 and the heater core 103 for generating a conditioning air.
The evaporator 141 serves as a cooling heat exchanger that evaporates a refrigerant circulating through the refrigerating cycle apparatus 100A and cools the air. The refrigerant circulating through the refrigerating cycle apparatus 100A is hereinafter referred to as the circulation refrigerant. The heater core 103 serves as a heating heat exchanger that heats the air using heat of a fluid, such as an engine coolant, flowing inside thereof.
An air mix door 104 is provided adjacent to the heater core 103. The air mix door 104 is configured to control a mixing ratio of air cooled by the evaporator 141 to air heated by the heater core 103. Thus, the temperature of air to be introduced into the passenger compartment 2 is controlled by means of the air mix door 104, thereby to control a temperature of the passenger compartment 2 to a desired temperature.
The refrigerating cycle apparatus 100A includes a compressor 110, a condenser 120, an expansion valve 131, and the evaporator 141. The compressor 110, the condenser 120, the expansion valve 131 and the evaporator 141 are connected in a due order through pipes 150. Thus, a closed circuit through which the circulation refrigerant flows is formed. Hoses, each having a rubber layer and/or a resin layer, are used on a suction side and a discharge side of the compressor 110.
Further, an internal heat exchanger 160 is provided to perform heat exchange between a high pressure circulation refrigerant between the condenser 120 and the expansion valve 13 and a low pressure circulation refrigerant between the evaporator 141 and the compressor 110.
The compressor 110 compresses the circulation refrigerant into a high temperature, high pressure condition. That is, the high pressure circulation refrigerant is generated by the compressor 110. The compressor 110 is driven by a driving force generated from an engine 10. A pulley 111 with an electromagnetic clutch is fixed to a driving shaft of the compressor 110. The driving force of the engine 10 is transmitted to the pulley 111 through a crank pulley 11 and a driving belt 12. The electromagnetic clutch is provided to intermittently connect or disconnect the pulley 111 and the driving shaft of the compressor 110. The compressor 110 is, for example, a variable displacement compressor.
The capacity of the compressor 110 is controlled by a control unit 105. A target pressure that is predetermined in accordance with a load is stored in the control unit 105. The control unit 105 controls the capacity of the compressor 110 such that an evaporation pressure of the evaporator 141 coincides with the target pressure, for example. For example, the control unit 105 controls the capacity of the compressor 110 such that the evaporation pressure is maintained in a range between equal to or greater than 0.2 MPa and equal to or greater than 0.3 MPa.
In the present embodiment, a surface temperature of the evaporator 141 is detected by a temperature sensor 106. The control unit 105 controls the capacity of the compressor 110 such that the surface temperature of the evaporator 141 is maintained to a target temperature.
Further, the control unit 105 controls the compressor 110 to conduct an intermittent operation in a low load region so as to avoid continuing a low capacity operation. The control unit 105 has a variable displacement control function for continuously controlling the capacity of the compressor 110 in an intermediate load region and a high load region and an intermittent control function for conducting an intermittent operation between a halted condition and a displacement condition to keep oil return in the low load condition. The intermittent operation is conducted by connecting and disconnecting the electromagnetic switch or by varying the capacity between a large region and a small region. As a result, the oil return is secured in the low load condition.
The condenser 120 serves as a high pressure-side heat exchanger. The condenser 120 is in communication with an outlet of the compressor 110. That is, the condenser 120 is disposed downstream of the compressor 110. The condenser 120 performs heat exchange between the circulation refrigerant and outside air, thereby to condense and liquefy the circulation refrigerant.
The expansion valve 131 serves as a decompressing device. The expansion valve 131 can be provided by a throttle, a valve, an ejector or the like. The expansion valve 131 isentropically decompresses a liquid phase circulation refrigerant flowing out from the condenser 120 to be expanded. The expansion valve 131 is arranged adjacent to the evaporator 141. The expansion valve 131 is, for example, a temperature sensing-type expansion valve that controls a throttle degree such that a condition of the refrigerant at a refrigerant outlet of the evaporator 141 is maintained to a predetermined condition.
The evaporator 141 serves as a low-pressure side heat exchanger. The evaporator 141 is also called a cooling device or a heat-absorbing device. The refrigerant outlet of the evaporator 141 is in communication with the suction side of the compressor 110 through the pipe 150 and the internal heat exchanger 160.
The internal heat exchanger 160 is exemplarily constructed of a double tube including an outer pipe 161 and an inner pipe 162 disposed inside of the outer pipe 161. The double tube 160 serves as a pipe. The double tube 160 extends generally in a front and rear direction over the engine compartment 1. Further, the double tube 160 is bent at plural locations to be appropriately arranged in the engine compartment 1.
The inner pipe 162 provides a low-pressure refrigerant passage therein to allow the low pressure circulation refrigerant, which has been decompressed by the expansion valve 131, to flow. A high-pressure refrigerant passage is provided between the inner pipe 162 and the outer pipe 161 to allow the high pressure circulation refrigerant to flow.
The double tube 160 has a predetermined length, such as at least 300 mm and at most 800 mm. A wall of the inner pipe 162 is formed with spiral grooves for enhancing heat exchange and for achieving a predetermined heat exchanging capacity within the predetermined length. The double tube 160 serves as the pipe. As such, the double tube 160 can be arranged while effectively using a limited space in the engine compartment 1. Further, the double tube 160 has a structure that is effective to reduce thermal influence by hot air in the engine compartment 1.
Next, the expansion valve 131 will be described with reference to FIG. 3. The expansion valve 131 has a block-shaped housing, and is thus generally called a box-type expansion valve. The expansion valve 131 has a valve portion 131 a for controlling the amount of refrigerant to be supplied to the evaporator 141 and a temperature sensing portion 131 b for controlling an opening degree of the valve portion 131 a.
The valve portion 131 a includes a valve seat, a valve body and a valve-closing spring. The temperature sensing portion 131 b includes a sensing member that senses a condition of the refrigerant at the outlet of the evaporator 141, a controlling member that generates an operation amount for operating the valve portion 131 a such that the condition of the refrigerant coincides with a target condition, and a driving member that controls the opening degree of the valve portion 131 a in accordance with the operation amount.
In the present embodiment, the temperature sensing portion 131 b includes a fluid pressure-type power element. The power element includes a diaphragm as a pressure sensing member. The diaphragm is arranged to separate a first chamber and a second chamber of the temperature sensing portion 131 b from each other. A rod member for driving the valve body is coupled to the diaphragm. As the diaphragm is displaced in response to a differential pressure between the first chamber and the second chamber, the opening degree of the valve portion 131 a is adjusted.
The evaporation pressure of the circulation refrigerant in the evaporator 141 is conducted to the first chamber. The second chamber is filled with a medium including a sealed two-phase refrigerant and a supplemental gas for adjustment. The sealed refrigerant has a saturated vapor pressure curve having a gradient greater than a gradient of a saturated vapor pressure curve of the circulation refrigerant.
The temperature of the refrigerant at the outlet of the evaporator 141 is transmitted to the medium filled in the second chamber. Thus, the sealed refrigerant senses the temperature of the refrigerant at the outlet of the evaporator 141. The sealed refrigerant varies the pressure of the second chamber in accordance with the temperature of the refrigerant at the outlet of the evaporator 141. Thus, the diaphragm is displaced in response to the difference between the evaporation pressure of the evaporator 141 and the pressure in accordance with the temperature of the circulation refrigerant at the outlet of the evaporator 141.
Next, an example of the circulation refrigerant will be described. In the present embodiment, a circulation refrigerant R0 is a mixture refrigerant. The circulation refrigerant R0 includes at least one component refrigerant having a GWP lower than a predetermined value. The circulation refrigerant R0 can be produced by mixing multiple refrigerants, which have been well-known. A property of the circulation refrigerant R0 is specified with respect to R134a refrigerant, which is a representative refrigerant widely used in a market, as a standard refrigerant.
Referring to FIG. 4, the property of the circulation refrigerant R0 will be described. FIG. 4 is generally called a p-h chart or a Mollier chart. In FIG. 4, a horizontal axis represents enthalpy H and a vertical axis represents pressure P. A solid line ML0 represents a saturation characteristic of the circulation refrigerant R0. A dashed line MLC represents a saturation characteristic of the standard refrigerant R134a.
A saturated liquid line of the saturation characteristic ML0 substantially coincides with a saturated liquid line of the saturation characteristic MLC. A critical pressure of the saturation characteristic ML0 is lower than a critical pressure of the saturation characteristic MLC. A saturated gas line of the saturation characteristic ML0 is located on a low enthalpy side of a saturated gas line of the saturation characteristic MLC. An enthalpy width of a two-phase region of the saturation characteristic ML0, which is defined between the saturated liquid line and the saturated gas line, is smaller than an enthalpy width of a two-phase region of the saturation characteristic MLC, as compared at an equal pressure.
At a pressure corresponding to a saturated temperature of 0 degree Celsius, an enthalpy width ED0 of the saturated characteristic ML0 is approximately 80% of an enthalpy width EDC of the saturated characteristic MLC. That is, the enthalpy width of the circulation refrigerant R0 is approximately −20% of the enthalpy width of the standard refrigerant R134a. Therefore, in a case where the circulation refrigerant R0 is used in a refrigerating cycle apparatus without having an internal heat exchanger, it is difficult to achieve a sufficient refrigerating capacity in an evaporating process.
In FIG. 4, a dashed chain line EL0 represents an isentropic line of the circulation refrigerant R0, and a double-dashed chain line ELC represents an isentropic line of the standard refrigerant R134a. The isentropic lines EL0, ELC each pass through a point of intersection between the saturated temperature of 0 degrees Celsius and the corresponding saturated gas line.
The isentropic line EL0 has a gradient greater than a gradient of the isentropic line ELC in a practical pressure range of a refrigerating cycle apparatus. The isentropic line EL0 has the gradient of 0.049 in a pressure range between 0.3 MPa and 2.0 MPa. The isentropic line ELC has the gradient of 0.040 in the pressure range between 0.3 MPa and 2.0 MPa. Here, the gradient is defined by P/H=MPa/(kJ/kg).
The compressor 110 compresses the refrigerant substantially along the isentropic line. Therefore, when the same compressor is used, the circulation refrigerant R0 achieves compressor efficiency higher than that of the standard refrigerant R134a. As such, in a case where the circulation refrigerant R0 is used, an increase in temperature in a compressing process is reduced and a discharge temperature, that is, a temperature of the refrigerant discharged from the compressor 110 is low, as compared with a case where the standard refrigerant R134a is used.
The refrigerating cycle apparatus 100A is operated to have a cycle behavior as shown by a solid line CY in FIG. 4. In the cycle CY, a condensing process is extended by an enthalpy width EH by means of the internal heat exchanger 160. Also, the evaporating process is extended by an enthalpy width EL by means of the internal heat exchanger 160. As such, an enthalpy width, which contributes to a cooling operation in the evaporator 141, is increased by the enthalpy width EH.
The enthalpy width EH, that is, the increase in the enthalpy width in the condensing process, corresponds to approximately 6% of the enthalpy width ED0 of the two-phase region of the circulation refrigerant R0. The internal heat exchanger 160 is configured to have a heat exchanging capacity so as to achieve the increase EH. In other words, the internal heat exchanger 160 is configured to increase the enthalpy width of the evaporating process by approximately 8% in the cycle CY under a high load idling operation condition. For example, the internal heat exchanger 160 has the heat exchanging capacity that can increase the temperature of the low pressure refrigerant in a range between equal to or greater than 5.0 degrees Celsius and equal to or less than 15.0 degrees Celsius.
Referring to FIG. 5, a valve-opening characteristic of the expansion valve 131 will be described. FIG. 5 is called a p-t chart. In FIG. 5, a horizontal axis represents temperature T and a vertical axis represents pressure P. A solid line SV0 represents a saturated vapor pressure curve of the circulation refrigerant R0. A dashed line SVC represents a saturated vapor pressure curve of the standard refrigerant R134a. A dashed chain line EV1 represents a control characteristic of the expansion valve 131. A double dashed chain line EV2 represents a control characteristic of a comparative example. The control characteristic EV1 of the expansion valve 131 is called a normal charge characteristic.
The normal charge characteristic is provided when the sealed refrigerant filled in the temperature sensing portion has a saturated vapor pressure curve same as or similar to the saturated vapor pressure curve of the circulation refrigerant R0. The control characteristic EV1 defines a curve that is substantially translated from the saturated vapor pressure curve SV0 in a direction parallel to the horizontal axis on the p-t chart.
The control characteristic EV1 provides the refrigerant at the outlet of the evaporator 141 with a superheat degree SH of approximately 5 degree Celsius, substantially over the entirety of an operation region of the refrigerating cycle apparatus 100A. FIG. 5 shows that the superheat degree SH of 5 degree Celsius is provided at an evaporation pressure where the temperature of the temperature sensing portion of the expansion valve 13 is approximately 0 degree Celsius.
Referring to FIG. 6, a relationship between a cooling capacity Q and a superheat degree SH will be described. The cooling capacity corresponds to a refrigerating capacity. In FIG. 6, a horizontal axis represents the superheat degree SH and a vertical axis represents the cooling capacity Q. In the present embodiment having the internal heat exchanger 160, the circulation refrigerant R0 exhibits the cooling capacity as shown by a capacity curve CP0. On the other hand, in a refrigerating cycle apparatus without having the internal heat exchanger 160, the circulation refrigerant R0 exhibits a cooling capacity as shown by a capacity curve CPC.
The capacity curve CP0 is higher than the capacity curve CPC. A difference between the cooling capacities CP0, CPC is caused by the following two reasons, for example. Firstly, the heat exchanging efficiency of the low pressure refrigerant in the internal heat exchanger 160 improves in a range where the superheat degree SH is equal to or greater than 0 degree Celsius. Secondly, an enthalpy standard temperature efficiency of the evaporator 141 improves in the range where the superheat degree SH is equal to or greater than 0 degree Celsius. Such a phenomenon is appreciated because a capacity Qea of the evaporator 141 is represented by the equation of Qea=φ·Gea·(ia−ir), in which φ denotes efficiency; Gea denotes an air volume; ia denotes inlet-side air enthalpy; and ir denotes saturated air enthalpy corresponding to the temperature of the refrigerant.
As shown in FIG. 6, for example, when the superheat degree SH is equal to or less than 7 degrees Celsius, a significant amount of improvement QD of the cooling capacity is achieved. As another example, an upper limit of the superheat degree SH can be set to 6 degrees Celsius or 5 degrees Celsius.
When the superheat degree SH is equal to or less than 7 degrees Celsius, oil can be stably returned to the compressor 110. Here, the stable oil return means that the oil return to the compressor 110 is observed within one minute in a wide operation region including the low load region, for example. The upper limit of the superheat degree SH can be set to 6 degrees Celsius or 5 degrees Celsius so as to ensure the oil return.
When the superheat degree SH is equal to or greater than 0 degrees Celsius, the expansion valve 131 is stably operated without hunting. Thus, the lower limit of the superheat degree SH can be set to 0 degrees Celsius. As another example, the lower limit of the superheat degree SH can be set to 0.5 degrees Celsius or 1.0 degree Celsius. As further another example, the lower limit of the superheat degree SH can be set higher than 1.0 degree Celsius where the peak of the cooling capacity Q is observed. As still another example, the lower limit of the superheat degree SH can be set to 2 degrees Celsius or 3 degrees Celsius.
The superheat degree SH can be controlled in the above discussed temperature ranges. For example, the superheat degree SH can be controlled in a range between equal to or greater than 0 degree Celsius and equal to or lower than 6 degrees Celsius. To achieve the further stable control, the superheat degree SH can be controlled to a higher side within the above range.
Next, an operation of the present embodiment will be described. As an air conditioning operation, such as a cooling operation, is commanded by a user, the compressor 110 is driven by the engine 10. Thus, the refrigerating cycle apparatus 100A begins operation. As a result, the temperature of the evaporator 141 reduces. The air blown by the blower 102 is cooled by the evaporator 141. The temperature of the air is further controlled to be introduced into the passenger compartment as the conditioned air.
During the operation of the refrigerating cycle apparatus 100A, the condition of the refrigerant at the outlet of the evaporator 141 is controlled by the expansion valve 131. The expansion valve 131 is operated such that the superheat degree SH of the refrigerant at the outlet of the evaporator 141 is maintained to approximately 5 degrees Celsius, for example. As a result, an improvement effect of the cooling capacity by the internal heat exchanger 160 is achieved. In this case, since the superheat degree SH is set to a relatively large degree, such as approximately 5 degrees Celsius, it is less likely that the opening degree of the expansion valve 131 will cause hunting. As such, the expansion valve 131 is stably operated.
The refrigerant flowing out from the evaporator 141 is further superheated by the internal heat exchanger 160. Therefore, a suction temperature, that is, the temperature of the refrigerant to be suctioned into the compressor 110 reaches a relatively high temperature. The superheated refrigerant is suctioned into the compressor 110 and compressed. Here, the isentropic line EL0 of the circulation refrigerant R0 has the gradient larger than the gradient of the isentropic line ELC of the standard refrigerant R134a. Therefore, an increase in temperature of the refrigerant inside of the compressor 110 is smaller in the case of the circulation refrigerant R0 than in the case of the standard refrigerant R134a. As a result, the discharge temperature in the case of the circulation refrigerant R0 can be substantially equal to or reduced lower than the discharge temperature in the case of the standard refrigerant R134a.
That is, it is less likely that the temperature of the refrigerant inside of and downstream of the compressor 110 will be excessively increased. As a result, it is less likely that the durability of rubber parts and resin parts, such as O-ring, hose, electronic devices and the like, used in an area from the suction portion to the discharge portion of the compressor 110, will be deteriorated due to heat.
During the operation of the refrigerating cycle apparatus 100A, the flow rate of the circulation refrigerant R0 is approximately 20% greater than the flow rate in the refrigerating cycle apparatus using the standard refrigerant R134a. The difference of the flow rates is caused because the circulation refrigerant R0 has a density higher than that of the standard refrigerant R134a under a low pressure vapor condition. In addition, the internal heat exchanger 160 increases the enthalpy width of the cycle CY. As a result, the refrigerating cycle apparatus 100A using the circulation refrigerant R0 exhibits the cooling capacity substantially similar to that of the refrigerating cycle apparatus using the standard refrigerant R134a.
Referring to FIG. 7, effects of the present embodiment will be described. FIG. 7 shows the cooling capacity Q and a discharge temperature TD of three refrigerating cycle apparatus, such as a refrigerating cycle apparatus using the standard refrigerant R134a without an internal heat exchanger, a refrigerating cycle apparatus using the circulation refrigerant R0 and without having an internal heat exchanger, and a refrigerating cycle apparatus using the circulation refrigerant R0 and having the internal heat exchanger 160 as the present embodiment. Also, the cooling capacity Q and the discharge temperature TD of an idling condition (IDLE) and a driving condition (DRIVE) of each refrigerating cycle apparatus are shown. The idling condition (IDLE) corresponds to a condition where the engine 10 is idling when a vehicle is at halt and the driving condition (DRIVE) corresponds to a condition where a vehicle is stably traveling. Further, the cooling capacity Q of the refrigerating cycle apparatus using the standard refrigerant R134a without having the internal heat exchanger is defined 100%.
The cooling capacity Q of the refrigerating cycle apparatus using the circulation refrigerant R0 without the internal heat exchanger is lower than the cooling capacity Q of the refrigerating cycle apparatus using the standard refrigerant R134a. The cooling capacity Q of the refrigerating cycle apparatus having the internal heat exchanger 160 and using the circulation refrigerant R0 as the present embodiment is slightly higher than the cooling capacity Q of the refrigerating cycle apparatus using the standard refrigerant R134a.
In addition, the circulation refrigerant R0 reduces the discharge temperature TD by about 10 degrees Celsius, as compared with the standard refrigerant R134a. Therefore, in the present embodiment, although the internal heat exchanger 160 is employed, the discharge temperature TD is substantially equal to or lower than that of standard refrigerant R134a.
In the present embodiment, the refrigerating cycle apparatus employs the circulation refrigerant R0, which has the property indicating the isentropic line EL0 having the gradient greater than the gradient of the isentropic line ELC of the standard refrigerant R134a as well as the saturation characteristic in which the enthalpy width of the two-phase region smaller than the enthalpy width of the two-phase region of the saturation characteristic of the standard refrigerant R134a on the p-t chart. Further, the refrigerating cycle apparatus employs the internal heat exchanger 160. Therefore, the cooling capacity substantially equal to that of the refrigerating cycle apparatus using the standard refrigerant R134a can be achieved while restricting the discharge temperature from excessively increasing. In addition, since the expansion valve 131 has the normal charge characteristic, the improvement effect of the cooling capacity by the internal heat exchanger 160 can be achieved in the entire operation region.
In the present embodiment, the refrigerant temperature can be reduced so that the durability of the parts is not deteriorated, while sufficiently achieving the refrigerating capacity by means of the internal heat exchanger 160. As such, selection of the refrigerant can be increased. Further, the refrigerating cycle apparatus has the advantage of using the circulation refrigerant R0 containing the component refrigerant having low GWP. As a result, the cooling capacity and the durability are improved while taking the advantage of the circulation refrigerant R0.

Second Embodiment

A second embodiment will be described with reference to FIG. 8. Hereinafter, parts similar to the first embodiment will be designated with like reference numerals, and features different from the first embodiment will be mainly described.
In the present embodiment, a refrigerating cycle apparatus 200A includes a compressor 210 driven by an electric motor, in place of the compressor 110 driven by the engine. Also, the refrigerating cycle apparatus 200A includes an expansion valve 231 having a control characteristic EV3, in place of the expansion valve 131 having the control characteristic EV1.
The compressor 210 includes a compressing section 211 and an electric motor 212. The compressing section 211 and the electric motor 212 are integrated with each other. For example, the compressing section 211 and the electric motor 212 are integrated so as to share a housing. The compressor 210 has a refrigerant suction port on a side of the electric motor 212. A control circuit 213 for driving the electric motor 212 is integrally mounted to the compressor 210.
The control circuit 213 is thermally connected to the compressor 210, and thus receives heat of the refrigerant. The control circuit 213 includes an inverter circuit. The control circuit 213 includes electronic circuit devices, such as a switching element, a photo-coupler, an IC element and the like.
The control circuit 213 includes an electronic circuit device having a relatively low heat resistance. For example, a photo-coupler has a heat resistance from approximately 100 degrees Celsius to approximately 110 degrees Celsius. Therefore, maintaining the temperature of the electronic circuit devices in a low temperature region is effective to restrict the deterioration of durability of the compressor 210.
The compressor 210 is formed with a refrigerant passage therein. The refrigerant passage passes through the inside of the electric motor 212 and reaches the compressing section 211. The refrigerant passage has a function of cooling the electric motor 212 and the control circuit 213. The compressor 210 is constructed so as to be adapted to the standard refrigerant R134a. Therefore, the compressor 210 can be employed in both a refrigerating cycle apparatus using the standard refrigerant R134a and a refrigerating cycle apparatus using the circulation refrigerant R0.
Referring to FIG. 9, the control characteristic EV3 of the expansion valve 231 will be described. The control characteristic EV3 is a normal charge characteristic that is set so that the superheat degree SH of the refrigerant at the outlet of the evaporator 141 is controlled to 2 degrees Celsius.
Next, an operation of the present embodiment will be described. As the electric motor 212 is driven, the circulation refrigerant R0 circulates through the refrigerating cycle apparatus 200A. The circulation refrigerant R0 cools the electric motor 212 and the control circuit 213 while flowing through the refrigerant passage of the electric motor 212.
While the refrigerating cycle apparatus 200A is in operation, the expansion valve 231 controls the superheat degree SH of the refrigerant at the outlet of the evaporator 141 to 2 degrees Celsius. The internal heat exchanger 160 further superheats the refrigerant.
In this case, since the superheat degree of the refrigerant at the outlet of the evaporator 141 is controlled to 2 degrees Celsius, the suction temperature of the compressor 210 can be relatively low, even though the internal heat exchanger 160 is employed. The expansion valve 231 controls the condition of the refrigerant at the outlet of the evaporator 141 such that the suction temperature at the suction port of the compressor 210 will not exceed the heat resistance temperature of the electronic circuit devices of the control circuit 213. For example, the expansion valve 231 controls the, superheat degree SH so that the suction temperature will not exceed the heat resistance temperature of the photo-coupler housed in the control circuit 213, such as about 100 and 110 degrees Celsius.
In the present embodiment, the refrigerating cycle apparatus 200A using the circulation refrigerant R0 employs the compressor 210, which can be also adapted to the standard refrigerant R134a. Further, the refrigerant temperature can be reduced to a relatively low temperature region so that the durability of the parts of the refrigerating cycle apparatus 200A will not be deteriorated, while sufficiently achieving the cooling capacity by means of the internal heat exchanger 160. As a result, the cooling capacity and the durability are improved while taking the advantage of the circulation refrigerant R0.
The control characteristic EV3 of the expansion valve 231 can be set such that the superheat degree SH coincides with a target temperature in a range between equal to or greater than 0 degree Celsius and equal to or less than 6 degrees Celsius. If the superheat degree SH exceeds 6 degrees Celsius, the refrigerant temperature will exceed the heat resistance temperature of the electronic circuit devices of the compressor 210. To satisfy both the high cooling capacity and the stable control, the lower limit of the superheat degree SH can be set to 1 degree Celsius.
Further, the upper limit of the superheat degree SH can be set to 3 degrees Celsius. The control characteristic EV3 of the expansion valve 231 can be set such that the superheat degree SH is controlled to 5 degrees Celsius.

Third Embodiment

Referring to FIG. 10, a third embodiment will be described. Parts similar to those of the first embodiment will be designated with like reference numerals, features different from the first embodiment will be mainly described.
In the present embodiment, a refrigerating cycle apparatus 300A includes an expansion valve 331 having a control characteristic EV4 shown in FIG. 11, in place of the expansion valve 131.
Referring to FIG. 11, the control characteristic EV4 of the expansion valve 331 will be described. The expansion valve 331 has a cross charge characteristic. The cross charge characteristic intersects the saturated vapor pressure curve SV0 of the circulation refrigerant R0. The cross charge characteristic creates a superheat degree in a high temperature region and a liquid-back condition in a low temperature region. The cross charge characteristic is provided since the saturated vapor pressure curve of the sealed refrigerant filled in the temperature sensing portion has the gradient smaller than the gradient of the saturated vapor pressure curve of the circulation refrigerant R0.
The control characteristic EV4 is set so that the superheat degree SH is controlled to 0 degrees Celsius when the saturated pressure P is approximately 0.2 MPaG and to 6 degrees Celsius when the saturated pressure P is approximately 0.3 MPaG. In this control characteristic, the superheat degree SH is achieved in a region where the saturated temperature T is higher than approximately 0 degrees Celsius. On the other hand, in a region where the saturated temperature T is lower than approximately 0 degrees Celsius, the superheat degree SH is 0 degrees Celsius. Thus, the liquid-back condition where large liquid component of the circulation refrigerant R0 is observed occurs.
The control characteristic EV4 is determined to intersect the saturated vapor pressure curve SV0 in the vicinity where the saturated temperature T is 0 degrees Celsius, considering that the refrigerating cycle apparatus 300A is used for cooling. In the cross charge characteristic, the hunting phenomenon of the valve opening degree is reduced, as compared with the normal charge characteristic. This is because the amount of change in the flow rate with respect to the amount of change in temperature is small in the cross charge characteristic, as compared with the normal charge characteristic.
In the present embodiment, the control characteristic EV4 of the expansion valve 331 is set so that the superheat degree SH for exhibiting high cooling capacity is implemented in a normal operation region where the high cooling capacity is required. In use for the cooling operation, the highest cooling performance is required when the evaporation temperature of the evaporator 141 is in a range between equal to or greater than 0 degrees Celsius and equal to or less than 10 degrees Celsius. This evaporation temperature corresponds to the saturated pressure P in a range between equal to or greater than approximately 0.2 MPaG and equal to or less than approximately 0.3 MPaG. As already described with reference to FIG. 6, the high cooling capacity is exhibited by controlling the superheat degree SH in the predetermined range. Therefore, in the present embodiment, the control characteristic EV4 is set so that the superheat degree SH is controlled in the range between equal to or greater than 0 degrees Celsius and equal to or less than 6 degrees Celsius, in the normal operation region where the saturated pressure P is equal to or greater than 0.2 MPaG and equal to or less than 0.3 MPaG.
In the present embodiment, the refrigerant temperature can be reduced to a relatively low temperature region so that the durability of the parts of the refrigerating cycle apparatus 300A will not be deteriorated, while sufficiently exhibiting the cooling capacity by the internal heat exchanger 160. As a result, the cooling capacity and the durability can be improved, while taking the advantage of the circulation refrigerant R0. In addition, since the cross charge characteristic is employed, the hunting phenomenon of the expansion valve 331 can be reduced.
The upper limit and the lower limit of the superheat degree SH provided by the cross charge characteristic EV4 can be set based on FIG. 6. For example, the upper limit and the lower limit of the superheat degree SH can be equal to or greater than 0 degrees Celsius and equal to or less than 5 degrees Celsius, respectively.

Other Embodiments

The present invention is not limited to the above described exemplary embodiments, but can be modified in various ways. For example, the exemplary embodiments will be modified as follows.
The isentropic line EL0 of the circulation refrigerant R0 can have a gradient in a range between equal to or greater than 0.044 and equal to or less than 0.054. By employing the refrigerant having the isentropic line with the gradient in the above range, even if the internal heat exchanger 160 is employed, the increase in the refrigerant temperature during the compressing process can be reduced. Thus, it is less likely that the discharge temperature will excessively rise. Accordingly, the durability of the part will not be deteriorated due to the heat.
The enthalpy width ED0 of the circulation refrigerant R0 can be set smaller than the enthalpy width EDC of the standard refrigerant R134a at least 15% and at most 25%. That is, the enthalpy width ED0 of the circulation refrigerant R0 can be set in a range between equal to or greater than 75% and equal to or less than 85% of the enthalpy width EDC of the standard refrigerant R134a. By employing the circulation refrigerant R0 having the above enthalpy width ED0, the cooling capacity substantially equal to that of the standard refrigerant R134a can be achieved as being supplemented by means of the internal heat exchanger 160.
The heat exchanging capacity of the internal heat exchanger 160 can be set so that the amount of increase EH of the enthalpy width by the internal heat exchanger 160 in the condensing process or in the evaporating process is in a range between equal to or greater than 5% and equal to or less than 10% of the enthalpy width ED0 of the circulation refrigerant R0 at the saturated temperature of 0 degree Celsius.
The above described exemplary embodiments can be adaptable to both the engine-driven compressor and the electric motor-driven compressor. When the electric motor-driven compressor is employed, the control characteristic of the expansion valve can be set such that the superheat degree SH is controlled to a relatively low value considering the heat resistance of the electronic circuit devices.
The circulation refrigerant R0 can be provided by a single component. Alternatively, the circulation refrigerant R0 can be various mixture refrigerant in which various components are contained with various ratios. Also, various refrigerants can be employed as the low GWP refrigerant contained in the circulation refrigerant R0.
The present invention can be employed to a refrigerating cycle apparatus including an ejector.
Additional advantages and modifications will readily occur to those skilled in the art. The invention in its broader term is therefore not limited to the specific details, representative apparatus, and illustrative examples shown and described.

Claims

1. A refrigerating cycle apparatus comprising:

a circulation refrigerant having a property that indicates an isentropic line and a saturation characteristic line on a p-h chart, the isentropic line having a gradient greater than an isentropic line of a R134a refrigerant, the saturation, characteristic line having a two-phase region enthalpy width smaller than a two-phase region enthalpy width of a saturation characteristic line of the R134a, the two-phase region enthalpy width being defined between a saturated liquid line and a saturated gas line;

a compressor that draws and compresses the circulation refrigerant into a high pressure circulation refrigerant;

a condenser-that condenses the high pressure circulation refrigerant;

an expansion valve that expands the high pressure circulation refrigerant into a low pressure circulation refrigerant;

an evaporator that evaporates the low pressure circulation refrigerant; and

an internal heat exchanger that performs heat exchange between the high pressure circulation refrigerant flowing from the condenser to the expansion valve and the low pressure circulation refrigerant flowing from the evaporator to the compressor.

2. The refrigerating cycle apparatus according to claim 1, wherein

at a saturated temperature of 0 degree Celsius on the p-h chart, the two-phase region enthalpy width of the circulation refrigerant is at most 85% of the two-phase region enthalpy width of the R134a refrigerant, and

the internal heat exchanger is configured to increase an enthalpy width of one of a condensing process and an evaporating process on the p-h chart by at least 5% of the two-phase region enthalpy width of the circulation refrigerant at the saturated temperature of 0 degree Celsius.

3. The refrigerating cycle apparatus according to claim 2, wherein

at the saturated temperature of 0 degrees Celsius on the p-h chart, the two-phase region enthalpy width of the circulation refrigerant is at least 75% and at most 85% of the two-phase region enthalpy width of the R134a refrigerant, and

the internal heat exchanger is configured to increase the enthalpy width of the one of the condensing process and the evaporating process on the p-h chart by at least 5% and at most 10% of the two-phase region enthalpy width of the circulation refrigerant at the saturated temperature of 0 degree Celsius.

4. The refrigerating cycle apparatus according to claim 1, wherein

the isentropic line of the circulation refrigerant passes through the saturated gas line of the circulation refrigerant at a saturated temperature of 0 degrees Celsius on the p-h chart, and

the gradient of the isentropic line of the circulation refrigerant is at least. 0.044 and at most 0.054.

5. The refrigerating cycle apparatus according to claim 1, wherein

the circulation refrigerant is a mixture refrigerant containing at least one low-GWP refrigerant.

6. The refrigerating cycle apparatus according to claim 1, wherein

the expansion valve has a normal charge characteristic to control a superheat degree of the circulation refrigerant at an outlet of the evaporator to a target value that is at least 0 degree Celsius and at most 6 degrees Celsius.

7. The refrigerating cycle apparatus according to claim 6, wherein

the compressor includes an electric motor and a control circuit, and

the control circuit is disposed to receive heat of the circulation refrigerant.

8. The refrigerating cycle apparatus according to claim 1, wherein

the expansion valve has a cross charge characteristic to control a superheat degree of the circulation refrigerant at an outlet of the evaporator to at least 0 degree Celsius and at most 6 degrees Celsius in a normal operation region.

9. The refrigerating cycle apparatus according to claim 8, wherein

the normal operation region corresponds to a range where refrigerant pressure at the outlet of the evaporator is at least 0.2 MPaG and at most 0.3 MPaG, and

the cross charge characteristic is set so that the superheat degree at the outlet of the evaporator is controlled to 0 degree Celsius when the refrigerant pressure at the outlet of the evaporator is 0.3 MPaG.

10. The refrigerating cycle apparatus according to claim 1, for being used for a vehicular air conditioning apparatus, wherein

the internal heat exchanger includes a double tube having an inner tube and an outer tube.