Refrigerating system with pressure control valve

Abstract: A refrigerating system of vapor compression type operating at a super critical area, while obtaining an increased efficiency. The refrigerating system includes a pressure control means (3) for controlling the temperature and the pressure at the outlet of a heat emitter (2). The pressure control valve (3) responds to a pressure difference between the inlet pressure of the refrigerant to the pressure control valve and the pressure in an outwardly sealed chamber (12) in which the refrigerant is filled such that, with respect to the volume of the chamber (12) under closed condition of the pressure control valve, a density of the refrigerant is in a range between a density of a saturated liquid at a temperature of 0°C and a density at the critical point of the refrigerant. As a result, the pressure and the temperature at the outlet of the heat emitter (2) is controlled substantially along the optimum control line ηmax, resulting in an effective execution of a refrigerating cycle at the critical area.

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a refrigerating system of a vapor compression type, provided with a pressure control valve for controlling a pressure of a refrigerant at an output side of a heat emitter. The present invention is suitably used for a refrigerating system where a carbon dioxide (CO 2 ), as a refrigerant, is used at a super critical region. 2. Description of Related Art Document NAKASHIMA Y., LIJIMA H., UMEHARA M., MATSUOKA F.: "Reversible-flow-type linear expansion valves for heat pumps" ASHRAE Trans (US), vol. 91 (part 2b), 1985, pages 1555-1568, discloses a refrigerating system of a vapor compression type comprising a compressor for compressing the refrigerant, a heat emitter for cooling the refrigerant compressed at the compressor, a pressure control valve for controlling the pressure of the refrigerant at the outlet of the heat emitter for reducing the pressure of the refrigerant and an evaporator for evaporating the refrigerant of the reduced pressure from the pressure control valve. Document EP-A-0 279 622 reveals a pressure control valve for use in a passageway for a refrigerant of a vapor compression type refrigerating system, said pressure control value comprising a partition wall for dividing the passageway into upstream and downstream chambers. Recently, it has been required to avoid the use of freon as a refrigerant in refrigerating systems. Thus, Japanese Unexamined Patent Publication No. 7-18602 discloses a vapor compression type refrigerating system where carbon dioxide (CO 2 ) is used as a refrigerant in place of freon. The carbon dioxide has a critical temperature as low as 31°C, which makes it possible for the temperature of the carbon dioxide at the outlet of the heat emitter to be higher than the critical temperature, so that the carbon dioxide is prevented from being condensed even at the outlet of the heat emitter. In other words, an increased difference in a enthalpy by condensing the refrigerant can not be obtained as is the case in the conventional refrigerating system using the freon. Thus, in order to increase an enthalpy difference, an increase in the pressure of the carbon dioxide at the outlet of the compressor, i.e., the pressure at the inlet of the heat emitter, is essential. However, an increase in the outlet pressure of the compressor may cause a coefficient of performance to be worsened. SUMMARY OF THE INVENTION An object of the present invention is to provide a refrigerating system using as, a refrigerant, carbon dioxide and operating at a critical region, capable of obtaining an increased efficiency. Another object of the present invention is to provide a refrigerating system using as, a refrigerant, carbon dioxide and operating at a critical region, capable of controlling the temperature and the pressure at an outlet of a heat emitter to obtain an operation substantially along an optimum operating line. According to the invention in claim 1, a deflector commences its deflection when the pressure at an upstream chamber is higher than the pressure in a sealingly closed chamber by a predetermined value, causing a valve member to open the valve port. Furthermore, a refrigerant is filled to a volume of said sealingly closed chamber as obtained when said valve member closes said valve port at a density in a range between a value of density of a saturated liquid state refrigerant of a temperature of 0°C and a value of density of a saturated liquid state refrigerant at the critical point. In this condition, the relationship between the pressure and the temperature of the refrigerant in the closed chamber substantially conforms to an optimum control line. Thus, when the pressure at the outlet of the heat emitter is increased to a pressure along the optimum control line, the pressure control valve opens its valve port. In other words, the outlet pressure and the outlet temperature of the heat emitter is controlled substantially along the optimum control line. Thus, an increased efficiency is obtained when executing the refrigerating cycle of a vapor compression type even at the supercritical range. BRIEF EXPLANATION OF ATTACHED DRAWINGS Fig. 1 is a schematic view of a refrigerating system with a pressure control valve according to a first embodiment of the present invention. Fig. 2 is a Mollier chart of a refrigerating cycle executed by the refrigerating system using carbon dioxide as the refrigerant. Fig. 3 is relationships between a pressure at the outlet of the heat emitter and coefficient of a performance (COP). Fig. 4 is a cross sectional view of the pressure control valve in Fig. 1. Fig. 5 is a Mollier chart, which is the same as Fig. 2 but shown with isodensity lines. Fig. 6 shows relationships between the volume of the closed chamber and an amount of carbon dioxide in the closed chamber. Fig. 7 is a cross sectional view of a pressure control valve in a second embodiment. Fig. 8 is a schematic view of a refrigerating system in a third embodiment. Fig. 9 is a cross sectional view of a pressure switch in a pressure control valve in Fig. 8. DESCRIPTION OF PREFERRED EMBODIMENTS First Embodiment Fig. 1 schematically illustrates a refrigerating system of a vapor compression type, which is constructed by a compressor 1, a heat emitter 2, a pressure control valve (expansion valve) 3, a heat absorber (evaporator) 4 and an accumulator 5, which are connected with each other by means of pipes 6 so that a closed circuit for a refrigerant (CO 2 ) is constructed. In a well known manner, the heat emitter 2 is arranged at a front portion of the vehicle, so that a large temperature difference is obtained between the outside air contacting the heat emitter and the carbon dioxide flowing in the heat emitter, thereby enhancing heat change efficiency. Furthermore, a fan 2-1 is arranged for generating a flow of the air for an effective emission of the heat from the heat emitter. The heat absorber 4 is arranged so as to be faced with a fan 4-1, which generates an air flow contacted with the heat absorber 4 and introduced into a cabin of a vehicle. As a result, a heat exchange is occurred between the carbon dioxide under a gas-liquid combined state and the air flow, causing the air flow to be cooled due to that fact that heat corresponding to a latent heat for evaporation is removed. At the accumulator 5, in the carbon dioxide, a separation of a gaseous phase is occurred, so that a liquid phase is stored in the accumulator 5. Furthermore, the compressor 1 is connected to a drive means such an internal combustion engine or an electric motor, so that a rotating movement is transmitted to the compressor. A refrigerating cycle as executed by the system using the carbon dioxide as the refrigerant in Fig. 1 will now be explained. Namely, in a Mollier chart in Fig. 2, an abscissa shows an enthalpy, while ordinate shows a pressure of the refrigerant. A curve SL is a saturated line on which the refrigerant is under a equilibrium condition. At the inlet of the compressor 1, the carbon dioxide is under a state designated by a point A in the Mollier chart in Fig. 2. Then, the gaseous state carbon dioxide is subjected to the compression at the compressor 1, so that the state of the refrigerant is moved to the one at a point B . The gaseous state carbon dioxide is subjected to a cooling at the heat emitter (gas cooler) 2, so that the emitted heat is taken by an air flow contacting the emitter, while the carbon dioxide is moved to a state designated by a point C in the Mollier chart. Then, the carbon dioxide is subjected to a pressure reduction at the pressure reducer 3, so that a gas-liquid combined state as designated by a point D in the Mollier chart is obtained. Finally, the gas-liquid combined state carbon dioxide is subjected to an evaporation at the evaporator 4, so that a flow of an air contacting with the evaporator 4 is cooled, while the carbon dioxide is moved to a gaseous state as designated by the point A for a repletion of the cycle. It should be noted that the phase change from the gaseous state to the gas-combined state occurs when the pressure of the carbon dioxide is reduced below the one on the saturating curve SL, i.e., the pressure at crossing point of the curve SL with the line CD. The carbon dioxide has, however, a critical temperature (a temperature at the critical point X) of about 31°C, which is lower than the critical temperature of freon, such as R12, of about 112°C. Thus, during a summer season, a situation may occur where the temperature of the carbon dioxide at the outlet of the heat emitter 2 is higher than its critical temperature, which prevents the carbon dioxide from being condensed. In other words, a crossing of the line BC with the saturation curve SL does not occur as shown in Fig. 2. In Fig. 2, curves with numerals are isothermal lines of temperature designated by the respective numerals. Furthermore, a condition at the outlet of the heat emitter 2 as designated by the point C is determined by the outlet pressure of the compressor 1 and the temperature of the carbon dioxide at the outlet of the heat emitter 2. The temperature of the carbon dioxide at the outlet of the heat emitter 2 is determined by the heat emission capacity at the heat emitter 2 and the temperature of the outside air. It is clear that the outside air temperature is not controllable, which makes the temperature of the carbon dioxide to be uncontrollable at the outlet of the heat emitter. In view of above, in order to control the condition at the outlet of the heat emitter designated by the point C in Fig. 2, a control of the outlet pressure of the compressor, which corresponds also to the pressure at the outlet of the heat emitter, is essential. Namely, in order to obtain a desired cooling capacity, i.e., an increased enthalpy difference under a condition of an increased outside air temperature in a summer season, it is essential that an outlet pressure of the compressor is increased to a value as designated by a point F in Fig. 2. In other words, the pressure at the outlet of the heat emitter 2 is increased to the value as designated by a point G in Fig. 2, so that the refrigerating cycle is executed in the order of E-F-G-H-E. In order to obtain an increased pressure at the outlet of the heat emitter 2, an increase in a pressure at the outlet of the compressor 1 is necessary, which necessitates an increased work done by the compressor, which corresponds to a change ΔL in the enthalpy during the compression process. However, such an increase in the enthalpy may cause a coefficient of performance COP to be reduced in a situation that the change ΔL in the enthalpy during the compression phase from the point A to the point B is larger than an increase Δi in the change in the enthalpy during the evaporation phase from the point D to A . The coefficient of performance COP is calculated by Δi/ΔL. A result of a calculation of the COP by using the pressure of the carbon dioxide at the outlet of the heat emitter will now be explained. A solid curve in Fig. 3 shows a relationship between the pressure at the outlet of the heat emitter 2 and the value of the COP when the temperature of the carbon dioxide is 40°C. The maximum value of the COP is obtained at a pressure P 1 of about 10 MPa. A dotted curve in Fig. 3 shows a similar relationship when the temperature of the carbon dioxide is 35°C. The maximum value of the COP is obtained at a pressure P 2 of about 8.0 MPa. The above process for obtaining the maximum value of the COP is done along the entire range of the temperature of the carbon dioxide. As a result, a curve ηmax, which is shown by a thick line in Fig. 2 and which is referred as the optimum control line hereinbelow is obtained, which makes COP the maximum value at designated conditions. In other words, in order to obtain an effective operation of the refrigerating system using carbon dioxide as the refrigerant, it is essential to provide a control means for controlling the pressure as well as the temperature of the carbon dioxide, so that they are located on the optimum control line ηmax at the outlet of the heat emitter 2. Note: The Mollier chart in Fig. 2 is extracted from Fundamentals Handbook, published by the American Society of Heating, Refrigerating and Air-conditioning Engineers. According to the present invention, the arrangement of the pressure control valve 3 is such that a desired control of the pressure of the carbon dioxide is done. Now, a construction of the pressure control valve 3 will be explained with reference to Fig. 4. The pressure control valve 3 includes a valve cover 10, a diaphragm 11 and a valve housing 13. The diaphragm 11 is, at its outer periphery, sandwiched between the cover 10 and the housing 13, so that a closed chamber 12 is formed above the diaphragm 11 and an opened chamber 12' is formed below the diaphragm. In the diaphragm chamber 12, the carbon dioxide is sealingly stored, while the volume of the carbon dioxide with respect to the volume of the chamber 12 as obtained at the seated state of the valve member 20 of the valve 1 is such that at a density of about 600 kg/m 3 of the carbon dioxide is obtained. The valve housing 13 is located in a section 6-1 of the recirculating pipe 6, so that a chamber 14 in communication with the heat emitter 2 for receiving the carbon dioxide therefrom is formed in the section 6-1 of the pipe 6 on the upper (upstream) side of the housing 13, while a chamber 15 in communication with the heat receiver 4 for issuing the carbon dioxide thereto is formed in the section 6-1 of the pipe 6 on the lower (downstream) side of the housing 13. The valve housing 13 is formed with a central bore 13-1 and a plate 16. The housing 6 is provided with a valve port 17 opened to the lower chamber 15, and diametrically opposite pair of communication holes 18 opened to the upper chamber 14. The pressure control valve 3 is further provided with a valve member 20, a valve spring 21 and a spacer 22. The valve member 20 has a bottom spherical end 20' facing the valve port 17 and a top end connected to the diaphragm 11 by a suitable conventional means, such as a nut (not shown). The spring 21 generates a spring force for urging the valve member 20 so that it is seated on the valve housing 13, thereby closing the valve port 17. The spacer 22 is for controlling a set force of the spring 21 to, for example, 1 MPa. Finally, an arrangement of the spacer 22 is such that the pressure at the central bore 13-1 in communication with the upper chamber 14 is opened to the lower diaphragm chamber 12'. As a result of this arrangement, a force is applied to the diaphragm 11 in accordance with the pressure difference between the upper and lower diaphragm chambers 12 and 12'. Namely, the pressure in the upper (closed) chamber 12 urges the diaphragm 11 to be moved downwardly, while the pressure in the lower chamber 12' opened to the chamber 14 urges the diaphragm 11 to be moved upwardly. Thus, the operation of the valve member 20 is controlled in accordance with the difference between the spring force and the force due to the pressure difference. Now, an operation of the pressure control valve 3 will be explained. Fig. 5 is a Mollier chart when the carbon dioxide (CO 2 ) is used and is same as that in Fig. 2. In Fig. 5, in addition to the isothermal lines with numerals designating respective temperatures of CO 2 , isodensity lines with numerals designating respective density of CO 2 are shown. In the closed upper diaphragm chamber 12, the carbon dioxide is sealingly stored at a density of 600 kg/m 3 when measured by the volume of the chamber 12 as obtained under the seated condition of the valve member to the valve port 17. Thus, the pressure and the temperature of the carbon dioxide in the sealed chamber 12 are varied along the isodensity line with the numeral of 600. When the temperature of CO 2 in the chamber 12 is, for example, 20°C, the pressure in the chamber 12 is 5.8 MPa, which is obtained as the pressure at the crossing point between the isothermal line of 20°C and isodensity line 600 kg/m 3 , which generates a force urging the valve member 20 to be seated on the valve seat. Furthermore, the set force of the spring 21 is about 1.0 MPa, which also generates a force urging the valve member 20 to be seated on the valve seat. In total, the valve member 20 is urged so as to be seated on the valve seat at the force of about 5.8 + 1.0 = 6.8 MPa. This means that, when the pressure inside the lower chamber 12', which urges the diaphragm 11 upwardly, i.e., the valve member 20 to be detached from the valve seat, is lower than 6.8 MPa, the valve member 20 maintains its closed position where it is seated on the valve seat, which prevents the carbon dioxide from flowing from the chamber 14 into the chamber 15. Contrary to this, when the pressure inside the chamber 12' is higher than 6.8 MPa, the diaphragm 11 is moved upwardly, which causes the valve member 20 to be detached from the valve seat, which allows the carbon dioxide to flow from the chamber 14 to the chamber 15 via the valve port 17. In the similar way, when the temperature inside the upper (closed) space 12 is 40°C, the pressure, which is the one at the crossing point between the isothermal line of 40°C and the isodensity line of 600 kg/m 3 , is 9.7 MPa. Thus, the total force for moving the valve member 20 toward the valve seat is 9.7 + 1 = 10.7 MPa. As a result, a pressure of the lower diaphragm chamber 12' lower than 10.7 MPa causes the valve plate 20 to be seated on the valve seat to close off the valve port 17, while the pressure of the lower diaphragm chamber 12' higher than 10.7 MPa causes the valve plate 20 to be detached from the valve seat to open the valve port 17. Now, an operation of the refrigerating system according to the present invention will be explained. When the system is under such a condition that, at the outlet of the heat emitter 2, the temperature is, for example, 40°C and the pressure is lower than 10.7 MPa, the pressure control valve 3 is under the closed condition where the valve member 20 is seated on the valve seat, which prevent CO 2 in the upstream chamber 14 from flowing into the downstream chamber 15. As a result, the CO 2 stored in the accumulator 5 is "sucked" by the compressor 1, which allows the pressure at the outlet of the heat emitter 2 to be increased from the one as obtained due to a state of a change of condition from point b' to c' to the one as obtained due to a state of change from point b" to c". When the pressure at the outlet of the heat emitter 2 exceeds 10.7 MPa, i.e., a change of the state from point B to C in Fig. 5 occurs, the diaphragm 11 is displaced upwardly and the valve member 20 is moved upwardly from the valve seat against the force of the spring 21, thereby allowing CO 2 in the upstream chamber 14 to flow into the downstream chamber 15. The CO 2 is subjected to an evaporation at the evaporator 4, while the air flow contacting the evaporator 4 is subjected to cooling. The CO 2 is returned into the accumulator 5, which causes the pressure to be reduced below 10.7 MPa at the outlet of the heat emitter 2, which causes the pressure control valve to take the closed position, again. In short, in the operation of the system, the closure of the pressure control valve causes the pressure at the outlet of the heat emitter 2 to be increased to a predetermined value, which is followed by the reduction of the pressure of CO 2 and the evaporation for cooling the air. In case where the temperature at the outlet of the heat emitter 2 is, for example, 20°C, the pressure control valve is under a closed condition when the pressure at the outlet of the heat emitter 2 is lower than about 6.8 MPa. When the pressure at the outlet of the heat emitter exceeds the value of about 6.8 MPa causes the pressure control valve to be opened. The pressure control valve in the embodiment in Fig. 4 has an operating characteristic such that it is opened when the pressure at the outlet of the heat emitter 2 is increased to a predetermined value. The operating characteristic is determined in accordance with the pressure characteristic at the sealed chamber 12 of the pressure control valve. As will be appreciated from Figs. 1 and 4, the isodensity curve at 600 kg/m 3 in the super critical area substantially conforms with the optimum control line ηmax. Thus, the pressure control valve in Fig. 4 operates to control the pressure along the optimum control line ηmax, so that an increased efficiency of execution of the refrigerating cycle is obtained by means of the refrigerating system. In the condensed area below the pressure at the critical point (X) in Fig. 2, a deviation of the isodensity line of 600 kg/m 3 from the optimum control line ηmax is increased. However, in the condensed area, the pressure inside the chamber 12 is located adjacent the saturated line SL. However, a set force applied to the valve member 20 by the coil spring 21 allows the system to be controlled to a super cooled condition (subcool condition) of about 10°C, which allows the system to be operated under a high efficiency even at a region of a pressure below the critical pressure. From the practical view point, it is desirable that the sealing of the CO 2 in the chamber 12 is such that the density of the CO 2 is in a range between the density of the saturated liquid CO 2 when the temperature is 0°C and the density of the saturated liquid of CO 2 when the temperature is that of CO 2 at the critical point. In particular, it is desirable that the density is in a range between the isodensity line D 2 of 450 kg/m 3 and the isodensity line D 1 of 950 kg/m 3 . This range of the density of CO 2 corresponds to the shaded area of the amount of CO 2 in Fig. 6, which shows relationships between the volume of the sealed chamber 12 and the amount of the sealed CO 2 . As will be understood from the operation of the pressure control valve, it is desirable that, in response to the temperature at the outlet of the heat emitter 2, i.e., the temperature in the space 14, the temperature in the chamber 12 varies without delay. Thus, it is desirable that the valve cover 10 and the valve housing 13 are made from a material of an increased heat conductivity for obtaining a conducted heat amount as large as possible. In view of this, according to the embodiment, the valve cover and the valve housing are made from brass, while the diaphragm 11, the valve member 20 and the coil spring 21 are made from a stainless steel. Finally, in the first embodiment, the valve cover 10 is advantageously formed with fins or a recess for increasing a heat transfer coefficient between the carbon dioxide in the space 14 and the valve cover 10. Second Embodiment Fig. 7 shows a modification of a pressure control valve 3 including a bellows member 23, in which the carbon dioxide is sealingly stored at a predetermined density. The bellows member is made of a brass material of an increased heat conductivity. A valve member 20 is connected to the bellows member 23 so that the valve member 20 is faced with a valve port 17 of the valve seat plate 16. A spring 21 for urging the valve member 20 toward the valve seat plate 16 is arranged in the sealed chamber 12 in the bellows member. An operation of the pressure control valve 3 in Fig. 7 is similar to that of the pressure control valve 3 in Fig. 4. Namely, a force due to the pressure of CO 2 stored in the chamber 12 plus the force of the spring 21 urges the valve member 20 to be seated on the valve seat 16. Contrary to this, the pressure of CO 2 in the conduit section 6 causes the valve member 20 to be detached from the valve seat 16. As a result, the pressure of the CO 2 in the conduit 6 is controlled to the pressure of the CO 2 in the chamber 12 plus the set force of the spring 21, as similar to the first embodiment in Fig. 4. Third Embodiment A third embodiment of the present invention is directed to a provision of a pressure control valve 3, which is electrically operated. Namely, as shown in Fig. 8, the pressure control valve 3 is constructed essentially by a sensor 24 for detecting a difference of the pressure of the recirculated carbon dioxide from the sealed reference pressure and an electrically operated expansion valve 25 which is operated by an electrically operating signal. It is desirable that the sensor section 24 is at a location on one side of the expansion valve 25 adjacent to the heat emitter 2. Fig. 9 shows a construction of the sensor part. Namely, the sensor section 24 includes a bellows 23 in which the carbon dioxide is sealingly stored at a predetermined density and a contact unit 26 constructed by a stationary contact 26-1 fixedly connected to the conduit 6 and a movable contact 26-2 connected to the bellows member 23 to face the stationary contact 26-1. The contact unit 26 is for controlling the operation of the expansion valve 25 via an amplifier (not shown). A pressure of CO 2 in the chamber 14 lower than the pressure in the sealed chamber 12 inside the bellows member 23 plus the set force of the spring 21 causes the bellows 23 to be inflated, thereby causing the contacts 26-1 and 26-2 to contact each other. As a result, the expansion valve 25 is controlled so that it is closed, thereby increasing the pressure at a location upstream from the pressure control valve 3. Contrary to this, a pressure of CO 2 at the location upstream from the pressure control valve higher than the pressure in the sealed chamber 12 inside the bellows member 23 plus the set force of the spring 21 causes the bellows 23 to be deflated, thereby causing the contacts 26-1 and 26-2 to be detached from each other. As a result, the expansion valve 25 is controlled so that it is opened, thereby reducing the pressure at the location upstream from the pressure control valve. In the above embodiments, the spring 21 is used for generating an initial set force. However, in place of the spring 21, an elastic force of the diaphragm or bellows can be used for generating the initial set force. Furthermore, in place of the coil spring, a non-condensable gas together with the refrigerant (CO 2 ) can be sealed in the space 12. Furthermore, an integral structure of the pressure control valve with the heat emitter can be employed.

Claims: A pressure control valve for use in a passageway for a refrigerant of a vapor compression type refrigerating system having, in addition to the pressure control valve (3), a compressor (1), a heat emitter (2), and an evaporator (4), said pressure control valve being arranged at a location between the heat emitter (2) and the evaporator (4), said pressure control valve comprising: a partition wall for dividing said passageway into an upstream chamber (14) for receiving the refrigerant from the heat emitter and a downstream chamber (15) for issuing the refrigerant to the evaporator; said partition wall defining a valve port (17) located between the upstream and downstream chambers; a deflector (11) having, therein, a sealingly closed chamber (12) and performing deflection in response to a pressure difference between said sealingly closed chamber (12) and the upstream chamber (14), and; a valve member (20) for controlling a communication of the upstream chamber (14) with the downstream chamber (15) via said valve port (17), said deflector commencing its deflection when a pressure at the upstream chamber is higher than a pressure in the sealingly closed chamber by a predetermined value; said deflector (11) being connected to said valve member (20) in such a manner that the commencing of the deflection of the deflector causes the valve member to be detached from the valve port; said sealingly closed chamber (12) being filled with a volume of said refrigerant which gives a density range of said refrigerant when said valve member (20) closes said valve port (17) of between a saturated liquid-state refrigerant at a temperature of 0° and a saturated liquid-state refrigerant at its critical point; and wherein said refrigerant is carbon dioxide and the density is in a range between 450 kg/m 3 and 950 kg/m 3 . A pressure control valve according to claim 1, wherein said deflector comprises a diaphragm (11) connected to the valve member (20), a casing (10) for storing therein the diaphragm so that the sealingly closed chamber (12) is formed on one side of the diaphragm, while the diaphragm being opened to the upstream chamber (14) on the side remote from the sealingly closed chamber and a spring (21) for urging the valve member so that the latter closes the valve port (17). A pressure control valve according to claim 1, wherein said deflector comprises a bellows (23) having a chamber therein as said sealingly closed chamber (12) and connected to said valve member, a support member for making the bellows to be stationary at the end opposite the valve member and a spring (21) for urging the valve member (20) so that the latter closes the valve port (17). A pressure control valve according to claim 1, wherein said valve member is of an electrically operated type, and wherein said deflector comprises a bellows member (23) defining therein a chamber as said sealingly closed chamber (12), a movable contact (26-2) connected to the bellows member, a stationary contact (26-1) faced with the movable contact, and a spring (21) for urging the contacts to be contacted with each other, thereby generating an electric signal for causing the valve member to close the valve port, the pressure of the upstream chamber higher than that in the sealingly closed chamber causing the movable contact to be detached from the stationary contact, causing the valve member to open the valve port. A refrigerating system comprising: a passageway for a refrigerant; a compressor (1) for compressing the refrigerant; a heat emitter (2) for cooling the refrigerant compressed at the compressor; a pressure control valve (3) for reducing the pressure of the refrigerant at an outlet of said heat emitter, and; an evaporator (4) for evaporating the refrigerant of the reduced pressure from the pressure control valve; wherein the refrigerating system comprises a pressure control valve according to claim 1.

Inventors: NISHIDA SHIN, ONIMARU SADAHISA, OZAKI YUKIKATSU, SAKAJO YUICHI, SAKAKIBARA HISAYOSHI, NISHIDA, SHIN, ONIMARU, SADAHISA, OZAKI, YUKIKATSU, SAKAJO, YUICHI, SAKAKIBARA, HISAYOSHI
Applicants: DENSO CORP, NIPPON SOKEN, DENSO CORPORATION, NIPPON SOKEN, INC.
Application year: 1997
Main classification: F25B9
IPC codes: F25B41, F25B1, F25B13, G05D16, C09K5, F25B9