JP2006161659A - Refrigerating cycle device - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a refrigerating cycle device having high performance and high reliability and capable of controlling the intermediate pressure of a two-stage compressor. <P>SOLUTION: The refrigerating cycle device having a compressor, which performs two-stage compression on a low-pressure side and a high-pressure side, has a refrigerant circuit for connecting a first heat exchanger for exchanging the heat between the refrigerant discharged from a high-pressure side compressing element of the compressor, a pressure reduction device for reducing the pressure of the refrigerant from the first heat exchanger and a second heat exchanger for exchanging the heat between the reduced-pressure refrigerant to each other through a refrigerant flow passage, and a refrigerant quantity control unit for controlling quantity of the refrigerant compressed by a low-pressure side compressing element and sucked to the high-pressure side compressing element. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a refrigeration cycle apparatus using a two-stage compressor.

  A heat pump apparatus used in a refrigeration cycle apparatus, for example, a refrigeration air conditioner or a hot water supply apparatus, generally includes a compressor, and there is one that uses a two-stage compressor as the compressor.

  The suction pressure and discharge pressure of the compressor used in such a refrigeration cycle apparatus vary depending on the operating conditions of the apparatus, and naturally the pressure ratio, which is the ratio of the discharge pressure and the suction pressure, also varies. In particular, when a two-stage compressor having a low pressure side compression element and a high pressure side compression element is used as the compressor, the suction pressure and the discharge pressure described above are the suction pressure (low pressure side suction pressure) of the low pressure side compression element and the high pressure side compression, respectively. This is the discharge pressure of the element (high-pressure discharge pressure).

  The discharge pressure of the low-pressure side compression element (low-pressure side discharge pressure) or the suction pressure of the high-pressure side compression element (high-pressure side suction pressure) of this two-stage compressor should be calculated based on the relationship and assumption described below. Can do.

  The relationship between the cylinder volume and the suction gas specific volume of each of the low pressure side compression element and the high pressure side compression element can be expressed as follows.

v _s1 = eta _v1・ V _th1・ N / G ₁ ……… (1)
v _s2 = eta _v2・ V _th2・ N / G ₂ ……… （2）
Where v _s is the suction gas specific volume, eta _v is the volumetric efficiency, V _th is the cylinder volume, N is the compressor rotational speed, G is the refrigerant mass flow rate, subscripts ₁ and ₂ are the low-pressure side compression element and high-pressure side compression, respectively. Indicates an element.

In addition, assuming that the compression of gas is an adiabatic change with “pv ^k = constant” (p: pressure, v: gas specific volume, k: adiabatic index), the discharge gas state of the low-pressure compression element and the suction of the high-pressure compression element Assuming that the gas state is the same, the following holds:

P _s1 v _s1 ^k = P _s2 v _s2 ^k (3)
Here, P _s1 and v _s1 indicate the suction pressure and suction gas specific volume of the low pressure side compression element, respectively, and P _s2 and v _s2 indicate the suction pressure and suction gas specific volume of the high pressure side compression element, respectively.

If v _s1 and v _s2 are eliminated from the above three equations, the following relationship holds between the high-pressure side suction pressure P _s2 and the low-pressure side suction pressure P _s1 .

P _s2 = P _s1 ((eta _v1 / eta _v2 ) ・ (V _th1 / V _th2 ) ・ (G ₂ / G ₁ )) ^k ……… (4)
Assuming that there is no inflow and outflow of the refrigerant with the outside between the low-pressure side outlet of the compression element and a suction port of the high pressure side compression element, the refrigerant mass in the refrigerant mass flow rate G ₁ and the high pressure side compression element of the low-pressure side compression element flow rate G ₂ is made the same, also when the volumetric efficiency eta _v2 volumetric efficiency eta _v1 and the high pressure side compression element of the low-pressure side compression element is assumed substantially the same as, the high-pressure side suction pressure P _s2 and the low-pressure side suction pressure P _s1 The relationship can be expressed as:

P _s2 ≒ P _s1 (V _th1 / V _th2 ) ^k (5)
That is, the high-pressure side suction pressure P _s2 is approximately (V _th1 / V _th2 ) ^k times the low-pressure side suction pressure P _s1 . Since V _th1 and V _th2 are cylinder volumes of the low-pressure side compression element and the high-pressure side compression element, (V _th1 / V _th2 ) is a constant value. Further, the change of the adiabatic index k due to the operating conditions is small. Therefore, the high-pressure side suction pressure P _s2 is _{approximately} a constant multiple of the low-pressure side suction pressure P _s1 . In the equation (5), when the high-pressure side suction pressure P _s2 is larger than the discharge pressure P _d (high-pressure side discharge pressure) of the compressor, P _s2 = P _d .

  Compared with a single-stage compressor, such a two-stage compressor reduces the pressure ratio and pressure difference at each stage, thereby reducing the load on each compression element and leakage in the working chamber, and improving efficiency. To do.

  The relationship between the ratio of the discharge pressure and the suction pressure of the low pressure side compression element (pressure ratio of the low pressure side compression element) and the ratio of the discharge pressure of the high pressure side compression element and the suction pressure (pressure ratio of the high pressure side compression element), or The relationship between the differential pressure between the discharge pressure of the low-pressure side compression element and the suction pressure (differential pressure of the low-pressure side compression element) and the differential pressure between the discharge pressure of the high-pressure side compression element and the suction pressure (differential pressure of the high-pressure side compression element) The efficiency can be improved by appropriately setting.

Japanese Patent Laid-Open No. 10-141270

However, the suction pressure (low pressure side suction pressure) and the discharge pressure (high pressure side discharge pressure) of the two-stage compressor change depending on the operating conditions, and the pressure ratio also changes. If the low-pressure side suction pressure P _s1 is determined, the high-pressure side suction pressure P _s2 is almost automatically determined by the equation (5).

  Therefore, depending on the operating conditions, the pressure ratio of the low-pressure side compression element and the pressure ratio of the high-pressure side compression element, or the relationship between the differential pressure of the low-pressure side compression element and the differential pressure of the high-pressure side compression element greatly deviates from the optimum point, There is a case where the original features of the two-stage compressor cannot be utilized.

  In particular, when the refrigeration cycle apparatus is operated at a low load and a low pressure ratio, the low-pressure discharge pressure (approximately equal to the high-pressure suction pressure) of the two-stage compressor reaches the compressor discharge pressure (high-pressure discharge pressure). Resulting in. Then, the compression work is not performed in the high-pressure side compression element, and the gas refrigerant simply circulates through the working chamber, so that the input energy becomes useless work. That is, there arises a problem that the efficiency is lower than that of the single stage compressor.

  In order to deal with this problem, in Patent Document 1, in order to control the effective volume of the compression chamber so that the substantial suction gas volume of the low-pressure side compression element decreases during low pressure ratio operation, the compression chamber of the low-pressure side compression element is used. And a suction bypass valve opening / closing device for opening and closing the suction bypass passage communicating with the suction side.

This is because the cylinder volume V _th1 of the low-pressure side compression element is reduced in the above-described equation (5) by operating the low-pressure ratio operation by the suction bypass passage provided in the low-pressure side compression element and the suction bypass valve opening / closing device that opens and closes it. This corresponds to the reduction of the high-pressure side suction pressure P _s2 .

  Therefore, it is possible to prevent the low-pressure side discharge pressure (approximately the same as the high-pressure side suction pressure) from approaching the discharge pressure. And the compression load in a low pressure side compression element and a high pressure side compression element approaches, and reduction of a compressor input is aimed at.

  However, Patent Document 1 has the following problems. In this prior art, a bypass hole is provided in a cylinder compression chamber of a low-pressure side compression element of a rolling piston type rotary two-stage compressor, and a suction bypass passage that connects the compression chamber and the suction side and a suction bypass valve that opens and closes the suction bypass passage are provided. Prepare.

  As the piston rotates in the cylinder, the suction chamber volume of the low-pressure compression element increases and performs a suction action, while the compression chamber volume decreases and performs a compression action and then a discharge action.

  During low pressure ratio operation where the pressure ratio is less than the set value, the suction bypass valve is opened and no substantial compression action is performed until the piston passes through the bypass hole in the cylinder compression chamber, and the refrigerant gas passes through the bypass passage. Thus, the refrigerant returns to the suction side of the low-pressure side compression element. And substantial compression action starts after a piston passes a bypass hole.

  Therefore, until the piston passes through the bypass hole in the cylinder compression chamber, the low-pressure side compression element does not perform the compression action but performs the pump action, and the refrigerant gas sent out from the compression chamber by the pump action returns to the suction side. Then, it is again sucked into the low pressure side compression element. Therefore, this section is performing useless work, and Patent Document 1 has a problem that the input of the compressor increases accordingly.

  In addition, since the rolling piston type rotary compressor performs a compression action and a discharge action while the piston rotates once in the cylinder, in the conventional technique disclosed in Patent Document 1, the compression action starts after the piston passes the bypass hole. Therefore, the rotation angle at which the refrigerant gas reaches the discharge pressure is also delayed, and therefore the discharge action must be performed between short rotation angles. Therefore, the flow rate of the refrigerant passing through the discharge port per unit time increases, which increases the discharge flow path loss, and there is a problem that the input of the compressor increases by this amount.

  Therefore, in the prior art disclosed in Patent Document 1, the low pressure side discharge pressure is prevented from approaching the high pressure side discharge pressure during low pressure ratio operation, and the compression loads in the low pressure side compression element and the high pressure side compression element are brought close to each other. Although the compressor input is reduced, the above problem has not been solved.

  An object of the present invention is to provide a high-performance and high-reliability refrigeration cycle apparatus including a two-stage compressor including a high-pressure side compression element that sucks refrigerant discharged from a low-pressure side compression element.

  To achieve the above object, a refrigeration cycle apparatus according to the present invention includes an electric element housed in a sealed container, a compressor including a low-pressure side compression element and a high-pressure side compression element that are driven by the electric element. A first heat exchanger that performs heat exchange of the refrigerant discharged from the high-pressure side compression element, a decompression device that lowers the pressure of the refrigerant from the first heat exchanger, and heat exchange of the decompressed refrigerant A refrigerant circuit that is connected to each of the second heat exchangers through a refrigerant flow path and sucks the refrigerant from the second heat exchanger into the low-pressure side compression element of the compressor; and the low-pressure side compression element And a refrigerant amount adjusting unit that adjusts the amount of refrigerant that is compressed by the high pressure side compression element and sucked into the high pressure side compression element.

  In order to achieve the above object, another refrigeration cycle apparatus of the present invention includes an electric element housed in a sealed container, a low-pressure side compression element and a high-pressure side compression element driven by the electric element. A compressor, a first heat exchanger that performs heat exchange of the refrigerant discharged from the high-pressure side compression element, and a first pressure reducing device that lowers the pressure of the refrigerant from the first heat exchanger; And a second heat exchanger for exchanging heat of the refrigerant connected to each other via a refrigerant flow path, and a refrigerant circuit for sucking the refrigerant from the second heat exchanger into the low-pressure side compression element of the compressor, An intermediate refrigerant flow path for communicating the refrigerant compressed by the low-pressure side compression element to a refrigerant flow path between the first pressure reducing device and the second heat exchanger; and a refrigerant from the intermediate refrigerant flow path; Second pressure reducing device in communication with the first or / and second pressure reducing device A control unit for adjusting the amount of the refrigerant is controlled to be sucked into the year the high pressure side compression element, and has a.

  In addition to the above configuration, the second decompression device may be provided on the intermediate refrigerant flow path.

  Further, in addition to the above configuration, the second decompression device is provided on the refrigerant flow path connected to the intermediate refrigerant flow path and downstream of the refrigerant flow from the connection portion with the intermediate refrigerant flow path. It may be.

  Further, in addition to the above configuration, the second pressure reducing device may be operated according to the suction pressure of the high pressure side compression element or the discharge pressure of the low pressure side compression element.

  In addition to the above configuration, the second pressure reducing device may operate according to a difference between a discharge temperature of the high-pressure side compression element and a discharge temperature of the low-pressure side compression element.

  In addition to the above configuration, the intermediate refrigerant flow path may further include a third heat exchanger for exchanging heat with the refrigerant compressed by the low pressure side compression element. Furthermore, the first heat exchanger and the third heat exchanger may have a configuration in which the refrigerant flowing through each of them exchanges heat with a common fluid.

  ADVANTAGE OF THE INVENTION According to this invention, the performance of the refrigerating-cycle apparatus provided with the two-stage compressor can be improved, and reliability can be improved.

  A refrigeration cycle apparatus according to the present invention includes an electric element housed in an airtight container, a compressor including a low-pressure side compression element and a high-pressure side compression element driven by the electric element, and a high-pressure side compression of the compressor A first heat exchanger for exchanging heat of the refrigerant discharged from the element, a decompression device for reducing the pressure of the refrigerant from the first heat exchanger, and a second heat exchange for exchanging heat with the decompressed refrigerant A refrigerant circuit that is connected to each other through a refrigerant flow path, and a refrigerant amount adjusting unit that adjusts the amount of refrigerant that is compressed by the low-pressure side compression element into the high-pressure side compression element. It is.

  The refrigeration cycle apparatus described with reference to FIG. 1 is a heat pump type water heater that is an embodiment of the present invention, and uses carbon dioxide as a refrigerant.

  In the apparatus of FIG. 1, the compressor 10 has two compression elements (a low pressure side compression element 11 and a high pressure side compression element 12) in an airtight container 13. The low pressure side compression element 11 and the high pressure side compression element 12 are driven by an electric element (not shown) housed in the hermetic container 13.

  The suction passage 11 a of the low-pressure side compression element 11 passes through the sealed container 13 and is connected to the outlet of the evaporator 32. The first discharge passage 11 b and the second discharge passage 11 b ′ of the low pressure side compression element 11 communicate with the inside of the sealed container 13. In other words, the first discharge passage 11 b and the second discharge passage 11 b ′ are open to the sealed container 13.

  The suction passage 12 a of the high-pressure side compression element 12 passes through the sealed container 13 and is connected to the second discharge passage 11 b ′ and the second refrigerant passage 22 of the heat exchanger 20. The discharge passage 12 b of the high pressure side compression element 12 is connected to the first refrigerant passage 21 of the heat exchanger 20.

  The flow of the gas refrigerant compressed by the low pressure side compression element 11 will be described. The gas refrigerant compressed by the low pressure side compression element 11 is discharged into the sealed container 13 through the first discharge passage 11b. The refrigerant in the sealed container 13 flows into the second discharge passage 11b '. At this time, the pressure in the sealed container 13 is equal to the discharge pressure of the low-pressure compression element 11.

  The refrigerant gas flowing through the second discharge passage 11b 'of the low-pressure side compression element 11 is branched into two flows at the connection portion with the suction passage 12a.

  One flow is sucked into the high-pressure compression element 12 via the suction passage 12 a of the high-pressure compression element 12. The gas refrigerant is further compressed by the high pressure side compression element 12 to become a high pressure gas refrigerant, and is discharged to the discharge passage 12b. The high-pressure gas refrigerant flows into the first refrigerant passage 21 of the heat exchanger 20 through the discharge passage 12b.

  The other flow flows into the second refrigerant passage 22 of the heat exchanger 20 through the suction passage 12a, and the intermediate refrigerant flow path where the divided refrigerant flows merge in the refrigerant flow path downstream of the first decompression device. It becomes.

  Next, the heat exchanger 20 will be described. The heat exchanger 20 includes a first refrigerant passage 21, a second refrigerant passage 22, and a water passage 23. The first refrigerant passage 21 and the water passage 23, and the second refrigerant passage 22 and the water passage 23 are provided with a contact structure so that heat exchange is possible.

  The inlet side of the first refrigerant passage 21 is connected to the discharge passage 12b, and the outlet side is connected to the first expansion valve 30 serving as a first pressure reducing device via a pipe line. The water passage 23 includes a first refrigerant passage 21 having a flow path structure that contacts the water passage 23 from the inlet portion to the outlet portion so as to allow heat exchange. Water is supplied to the water passage 23 in the direction of the arrow in FIG. That is, the flow of water is directed in a direction opposite to the flow of the refrigerant.

  Compared to the first refrigerant passage 21, the second refrigerant passage 22 has a flow path structure that does not exchange heat with the outlet side portion of the water passage 23. The second refrigerant passage 22 communicates with the second discharge passage 11b ′ of the low-pressure compression element 11 through the suction passage 12a on the inlet side, and the second expansion passage 31 serving as the second decompression device on the outlet side of the second refrigerant passage 22 Connect through the road. The gas refrigerant flowing through the second refrigerant passage 22 exchanges heat with the water flowing through the water passage 23 and flows to the second expansion valve 31.

  In the present embodiment, the first refrigerant passage 21 of the heat exchanger 20 communicates with the first expansion valve 30, and the second refrigerant passage 22 communicates with the second expansion valve 31. After the refrigerant that has passed through the first expansion valve 30 and the refrigerant that has passed through the second expansion valve 31 are mixed, the mixed refrigerant flows into the evaporator 32. If there is no refrigerant passing through the second expansion valve 31, only the refrigerant that has passed through the first expansion valve 30 flows into the evaporator 32. Details of the flow of the refrigerant according to the opening of each expansion valve will be described later.

  The refrigerant having exchanged heat with the evaporator 32 is sucked into the low-pressure compression element 11 through the suction passage 11a. The blower 33 blows air to the evaporator 32, and the evaporator 32 performs heat exchange between the refrigerant and the air.

  Next, the control system in FIG. 1 will be described. First, the first control device 50 is a control device that controls the degree of superheat in the evaporator 32, and the second control device 51 controls the refrigerant gas discharge temperature difference from the high pressure side compression element 12 and the low pressure side compression element 11. It is a control device.

Evaporation temperature sensor 40 is a refrigerant heat transfer tubes of the evaporator 32 (not shown) provided in contact with the surface, approximately detecting the evaporation temperature T _e of the refrigerant. Further, the low-pressure side suction temperature sensor 41 is provided in contact with the pipe surface, which is the suction passage 11a of the low-pressure side compression element 11 of the compressor 10, and approximates the suction gas refrigerant temperature T _s1 of the low-pressure side compression element 11. Detect.

The low-pressure side discharge temperature sensor 42 is provided in contact with the pipe surface which is the discharge passage 11b ′ of the low-pressure side compression element 11, and approximately detects the discharge gas refrigerant temperature _Td1 of the low-pressure side compression element 11. Further, the high pressure side discharge temperature sensor 43 is provided in contact with the pipe surface which is the discharge passage 12b of the high pressure side compression element 12, and detects the discharge gas refrigerant temperature _Td2 of the high pressure side compression element 12 approximately.

The first controller 50 the amount of superheat is the difference between the evaporation temperature T _e detected by the intake gas temperature T _s1 and the temperature sensor 40 detected by the temperature sensor 41, the opening degree of the first expansion valve 30 Control. For example, when the target superheat degree is set to 5 ° C., the first controller 50 controls the degree of opening and closing of the first expansion valve 30 so that the superheat degree becomes the target superheat degree.

The second controller 51 detects the difference between the high-pressure side and low-pressure side discharge temperatures (T _d2 −), which is the difference between the high-pressure side discharge temperature T _d2 detected by the temperature sensor 43 and the low-pressure side discharge temperature T _d1 detected by the temperature sensor 42. The magnitude of T _d1 ) is controlled by the opening of the second expansion valve 31. For example, as shown in FIG. 2, the target discharge temperature difference (T _d2 −T _d1 ) is set as a linear function of the difference between the high pressure side discharge temperature T _d2 and the low pressure side suction temperature T _s1 (T _d2 −T _s1 ). The second controller 51 controls the valve opening / closing degree of the second expansion valve 31 so that the temperature difference becomes the target value.

  In the present embodiment, carbon dioxide is used as the refrigerant of the heat pump type hot water heater, and the cylinder volume of the high pressure side compression element 12 is set to 60% of the cylinder volume of the low pressure side compression element.

  The operation of the refrigeration cycle during the high pressure ratio operation and the low pressure ratio operation of the heat pump type water heater configured as described above will be described.

FIG. 3 is an explanatory diagram of the refrigerant flow when the two-stage compressor of the heat pump type hot water heater performs a high pressure ratio operation. For example, the suction pressure, discharge pressure and pressure ratio as winter pressure conditions are the suction pressure (low-pressure side suction pressure) P _s1 = 3.5 MPa, the discharge pressure (high-pressure side discharge pressure) P _d = 10 MPa, and the pressure ratio P _d / _Let P _s1 = 2.9.

At this time, the second expansion valve 31 is controlled so that the discharge temperature difference (T _d2 −T _d1 ) between the high pressure side compression element 12 and the low pressure side compression element 11 becomes the target value shown in FIG. Fully closed state. In this case, the refrigerant flow rate G ₂ of the refrigerant flow rate G ₁ and the high pressure side compression element 12 of the low pressure side compression element 11 are the same.

Further, assuming that the volumetric efficiency eta _v1 of the low pressure side compression element 11 and the volumetric efficiency eta _v2 of the high pressure side compression element 12 are substantially the same, the cylinder volume ratio V _th2 / V _th1 = 0.6 of the high pressure side compression element and the low pressure side compression element. When the adiabatic index k = 1.3 is used, the high-pressure side suction pressure is P _s2 = 6.8 MPa from the above equation (4). At this time, the pressure ratio P _s2 / P _{s1 of} the low pressure side compression element 11 is 1.9, the pressure ratio P _d / P _{s2 of} the high pressure side compression element 12 is 1.5, and the pressure difference P _s2 −P _{s1 of the} low pressure side compression element 11 = 3.3 MPa, and the pressure difference P _d −P _{s2 of} the high pressure side compression element 12 is 3.2 MPa.

  Therefore, the pressure ratio and pressure difference of each compression element approach, the load of each compression element and the leakage of the working chamber are reduced, and the efficiency of the compressor is improved.

FIG. 4 is an explanatory view of the refrigerant flow when the two-stage compressor of the heat pump type hot water heater of FIG. 1 performs the low pressure ratio operation. For example, the suction pressure, discharge pressure, and pressure ratio for summer pressure conditions are suction pressure (low pressure suction pressure) P _s1 = 5.6 MPa, discharge pressure (high pressure discharge pressure) P _d = 10 MPa, and pressure ratio P _d / P _s1 = 1.8.

At this time, the second expansion valve 31 is controlled so that the discharge temperature difference (T _d2 −T _d1 ) between the high pressure side compression element 13 and the low pressure side compression element 11 becomes the target value shown in FIG. About 25% of the refrigerant flowing from the low pressure side compression element 11 to the high pressure side compression element 12 is circulated through the second refrigerant passage of the heat exchanger 20. Therefore, the refrigerant flow ratio G ₂ / G ₁ = 0.75.

Also, assuming a volumetric efficiency ratio of eta _v1 / eta _v2 = 1.0, and using a cylinder volume ratio V _th2 / V _th1 = 0.6 and an adiabatic index k = 1.3, the high-pressure side suction pressure is P _s2 = 7.5MPa. At this time, the pressure ratio P _s2 / P _s1 = 1.3 of the low pressure side compression element 11, the pressure ratio P _d / P _s2 = 1.3 of the high pressure side compression element 12, and the pressure difference P _s2 −P _{s1 of the} low pressure side compression element 11 = 1.9 MPa, the pressure difference P _d −P _{s2 of} the high pressure side compression element 12 is 2.5 MPa.

  Therefore, the pressure ratio and pressure difference of each compression element approach, the load of each compression element and the leakage of the working chamber are reduced, and the efficiency of the compressor is improved.

Here, at the time of low pressure ratio operation, when there is no means for branching the refrigerant in the discharge passage 11b ′ of the low pressure side compression element 11, the refrigerant flow rate ratio G ₂ / G ₁ = 1.0. The suction pressure is P _s2 = 10.9 MPa.

Since this value is higher than the discharge pressure P _d = 10 MPa, P _s2 = P _d . Therefore, only the low pressure side compression element 11 reaches the discharge pressure of the compressor, and the high pressure side compression element 12 does not perform compression work. In other words, the gas refrigerant simply circulates in the working chamber, and is performing a useless work, and the efficiency is lower than that of the single-stage compressor.

  Therefore, this embodiment can solve this problem. Further, in the prior art in which the refrigerant in the compression chamber of the low-pressure side compression element is bypassed to the suction side, the compressor requires an extra pump work for bypassing the refrigerant in addition to the compression work, whereas the present invention provides a pump work. It is more efficient because it does not require extra power.

  Further, as in the present embodiment, since the refrigerant flowing into the second refrigerant passage 22 of the heat exchanger 20 is the discharge gas after the low-pressure side compression, it flows into the first refrigerant passage 21 that is the discharge gas after the high-pressure side compression. The temperature is lower than that of the refrigerant.

For example, assuming adiabatic compression with low-pressure side suction pressure P _s1 = 5.6 MPa, low-pressure side suction temperature T _s1 = 24 ° C, high-pressure side suction pressure P _s2 = 7.5 MPa, discharge pressure P _d = 10 MP, low-pressure side discharge temperature T _d1 = 46 ° C, high-pressure side discharge temperature T _d2 = 68 ° C.

  The first refrigerant passage 21 of the heat exchanger 20 is disposed so as to face the entire flow from the inlet to the outlet of the water passage 23, whereas the second refrigerant passage 22 extends from the inlet of the water passage 23 to the middle. It arrange | positions so that it may become a counterflow.

  Therefore, the temperature of the high-temperature refrigerant (for example, 68 ° C.) that has flowed into the first refrigerant passage 21 decreases while heating the water flowing in the water passage 23, and the second refrigerant passage 22 is near the inlet of the second refrigerant passage 22. It becomes equivalent to the refrigerant temperature at the inlet (for example, 46 ° C.). Thereafter, the refrigerant in the first refrigerant passage 21 and the refrigerant in the second refrigerant passage 22 heat the water in the water passage 23 at the same temperature level, and the temperature similarly decreases. Therefore, since the second refrigerant passage is arranged so that the refrigerant temperatures in the two refrigerant passages 21 and 22 at the same position are at the same level, heat exchange between the refrigerant and water can be performed efficiently.

  With the above configuration and operation, the suction pressure of the high-pressure side compression element, which is the intermediate pressure of the two-stage compressor, can be appropriately controlled, and high-efficiency operation can be achieved regardless of the pressure ratio.

  The configuration of FIG. 5 is obtained by adding a pressure sensor 44 that detects the pressure in the compressor hermetic container 13 to the configuration of FIG. 5, parts that are the same as those in FIG. 1 are given the same reference numerals, and descriptions thereof are omitted.

  Carbon dioxide used as a refrigerant has a critical point temperature as low as 31 ° C., and when applied to a refrigeration cycle of a heat pump device and a refrigeration air conditioner, it becomes a transcritical refrigeration cycle crossing the critical point.

  In the supercritical state, elution of a polymer material such as a resin becomes a problem, and when water is present, elution is further promoted and there is a possibility that the resin material is significantly deteriorated. An electric element (not shown) for driving the compression elements 11 and 12 is accommodated in the sealed container 13 of the compressor 10, and a resin material is used for the electric element. Therefore, in the present embodiment, the resin material of the electric element is prevented from being deteriorated by keeping the pressure in the sealed container 13 below the critical pressure.

The critical pressure of carbon dioxide is about 7.4MPa. During operation of the heat pump type water heater, the pressure sensor 44 detects the pressure in the hermetic container of the compressor. For the suction pressure (low pressure suction pressure), which is a low pressure of the compressor, and the discharge pressure (high pressure discharge pressure), which is a high pressure, the pressure in the sealed container is an intermediate pressure equivalent to the low pressure discharge pressure and the high pressure suction pressure. It has become. The second controller 51 has an upper limit pressure where the intermediate pressure detected by the pressure sensor 44 is lower than the critical pressure, for example, 7.0 MPa or less, and the discharge temperature difference (T _d2 −T _d1 ) between the high pressure side and the low pressure side is the above figure. The second expansion valve 31 is controlled so that the target value shown in FIG.

Under the same high pressure ratio operation conditions as described in FIG. 3, the second expansion valve 31 is fully closed when the suction pressure (low-pressure side suction pressure) P _s1 = 3.5 MPa and the discharge pressure (high-pressure side discharge pressure) P _d = 10 MPa. Thus, the high-pressure side suction pressure (equivalent to the pressure in the sealed container) is P _s2 = 6.8 MPa, which is the same as in Example 1 and is lower than the upper limit pressure 7.0 MPa.

On the other hand, when the suction pressure (low-pressure side suction pressure) P _s1 = 5.6 MPa and discharge pressure (high-pressure side discharge pressure) P _d = 10 MPa under the same low pressure ratio operation conditions as described in FIG. When the temperature difference (T _d2 −T _d1 ) control is performed, the intermediate pressure becomes 7.5 MPa as in the first embodiment and exceeds the upper limit of 7.0 MPa. Therefore, the intermediate pressure detected by the pressure sensor 44 is higher than the upper limit of 7.0 MPa. The expansion valve 31 is controlled so that At this time, when calculated from the equation (4), approximately 29% of the refrigerant flowing from the low pressure side compression element 11 to the high pressure side compression element 12 is circulated through the second refrigerant passage of the heat exchanger 20.

  Therefore, since the pressure in the sealed container 13 of the compressor 10 can be kept below the critical pressure, deterioration due to supercritical elution of the resin material such as an electric element housed in the sealed container 13 is prevented, and reliability is improved. Can be improved.

  Moreover, since the pressure in the sealed container of the compressor is suppressed to a critical pressure or less, the pressure resistance of the sealed container can be lowered, and the cost can be reduced.

  The configuration of FIG. 6 is obtained by adding an intermediate cooler of a two-stage compressor to the configuration of FIG. 6, parts that are the same as those in FIG. 1 are given the same reference numerals, and descriptions thereof are omitted.

  In the heat exchanger 20 ′, the first refrigerant passage 21 is disposed so as to face the entire flow from the inlet to the outlet of the water passage 23.

  The inlet of the second refrigerant passage intermediate cooling portion 24 a is connected to the second discharge passage 11 b ′ of the low pressure side compression element 11 of the compressor 10. Moreover, the intermediate cooling part 24a is located in the intermediate part of the water passage 23, and is arrange | positioned so that heat exchange with the water used as a counterflow is possible.

  Further, the second refrigerant passage bypass cooling part 24b is arranged on the inlet side of the water passage 23 so as to be able to exchange heat with the counterflow. One of the two branches at the outlet of the second refrigerant passage intermediate cooling portion 24a is connected to the suction passage 12a of the high-pressure side compression element 12, and the other is connected to the inlet side of the second refrigerant passage bypass cooling portion 24b. . Furthermore, the outlet of the second refrigerant passage bypass cooling unit 24 b is connected to the inlet of the evaporator 32 via the second expansion valve 31.

  At this time, the gas refrigerant compressed from the low pressure to the intermediate pressure by the low pressure side compression element 11 of the compressor 10 and the temperature rising is stored in the water passage 23 by the second refrigerant passage intermediate cooling portion 24a of the heat exchanger 20 ′. It is cooled by exchanging heat with water. Therefore, since the refrigerant temperature at the inlet of the high pressure side compression element 12 is lowered, the theoretical adiabatic compression work per unit mass is reduced, and the power consumption of the heat pump type hot water heater can be reduced.

  Further, as described in detail with reference to FIG. 1, by controlling the second expansion valve 31, the intermediate pressure of the two-stage compressor is appropriately controlled, so that highly efficient operation can be performed regardless of the operation pressure ratio. Can be planned.

  FIG. 7 is a refrigeration cycle configuration diagram illustrating an outline of a heat pump heater according to an embodiment of the present invention. The configuration is changed from the water-refrigerant heat exchanger in the configuration of the embodiment of FIG. 1 to the air-refrigerant heat exchanger. 7, parts that are the same as those in FIG. 1 are given the same reference numerals, and descriptions thereof are omitted.

  The inlet of the main gas cooler 60 is connected to the discharge passage 12 b of the high pressure side compression element 12 of the compressor 10, and the outlet of the main gas cooler 60 is connected to the first expansion valve 30. Further, the inlet of the intermediate cooler 61 is connected to the second discharge passage 11 b ′ of the low pressure side compression element 11 of the compressor 10, the outlet of the intermediate cooler 61 is branched into two, one side of the high pressure side compression element 12. The other side is connected to the inlet of the bypass gas cooler 62. Furthermore, the outlet of the bypass gas cooler 62 is connected to the inlet of the evaporator 32 via the second expansion valve 31.

  The coolers 60, 61, 62 are arranged in the order of relatively high refrigerant temperatures, and the gas refrigerant can be efficiently exchanged heat so that it becomes a pseudo counter flow with the air blown by the blower 63. it can.

  Further, as described in detail with reference to FIG. 1, by controlling the second expansion valve 31, the intermediate pressure of the two-stage compressor can be appropriately controlled, and the highly efficient operation is performed regardless of the operation pressure ratio. Can be achieved.

  Furthermore, similarly to the embodiment of FIG. 6, the power consumption of the heat pump heater can be reduced by performing the intermediate cooler of the two-stage compressor.

  Although the example which applied one Embodiment of this invention to the heat pump type water heater and the heater was described as a refrigeration cycle apparatus using a two-stage compressor, it can be similarly applied to a refrigeration apparatus using a two-stage compressor. .

  In addition, the internal heat exchanger that exchanges heat between the refrigerant in the suction passage of the compressor and the refrigerant in the passage before the expansion valve has not been described, but for the purpose of improving the efficiency of the device and increasing the heating temperature of water or air, An internal heat exchanger may be provided.

  The configuration of FIG. 8 is obtained by applying another heat exchanger applicable as the heat exchanger of FIG. In the configuration of FIG. 8, the same components as those in FIG.

  The heat exchanger 20 in FIG. 8 will be described. The heat exchanger 20 includes a first refrigerant passage 21 including a plurality of pipes, a second refrigerant passage 22 and a water passage 23.

  9, the first refrigerant passage 21 and the water passage 23, and the second refrigerant passage 22 and the water passage 23 have a contact structure so that heat exchange is possible.

  The first refrigerant passage 21 and the discharge passage 12b are connected via a distributor (not shown in detail) provided at the inlet of the first refrigerant passage 21. The exit portion of the first refrigerant passage 21 is a conduit that joins a plurality of pipes. The outlet of the first refrigerant passage 21 where the pipes merge together is connected to the first expansion valve 30 that is a first decompression device.

  Compared to the first refrigerant passage 21, the second refrigerant passage 22 has a flow path structure that does not exchange heat with the outlet side portion of the water passage 23. The first refrigerant passage 21 communicates with the first expansion valve 30, and the first expansion valve 30 communicates with a second expansion valve 31 that is a second decompression device.

  The outlet of the second refrigerant passage 22 is connected to a pipe between the first expander 30 and the second expander 31 that is the second decompression device.

  Therefore, if there is a refrigerant that passes through the second refrigerant passage 22, it mixes with the refrigerant that has passed through the first expansion valve 30, and then flows into the evaporator 32 via the second expansion valve 31.

  The operation of the refrigeration cycle in the high pressure ratio operation and the low pressure ratio operation in the heat pump type hot water heater of FIG. 8 will be described.

For example, as the pressure conditions when operating in winter, as in the description of FIG. 3, the suction pressure (low pressure suction pressure) P _s1 = 3.5 MPa and the discharge pressure (high pressure discharge pressure) P _d = 10 MPa. The pressure ratio P _d / P _s1 = 2.9. At this time, the second expansion valve 31 is controlled so that the discharge temperature difference (T _d2 −T _d1 ) between the high pressure side and the low pressure side becomes the target value shown in FIG. At this time, the pressure in the compressor 13 and the pressure at the outlet of the first expansion valve 30 are substantially the same, and no refrigerant flows through the second refrigerant flow path 22.

Thus, refrigerant flow rate G ₂ of the refrigerant flow rate G ₁ and the high pressure side compression element 12 of the low pressure side compression element 11 are the same. Further, assuming that the volumetric efficiency eta _v1 of the low pressure side compression element 11 and the volumetric efficiency eta _v2 of the high pressure side compression element 12 are substantially the same, the cylinder volume ratio V _th2 / V _th1 = 0.6 of the high pressure side compression element and the low pressure side compression element. When the adiabatic index k = 1.3 is used, the high-pressure side suction pressure is P _s2 = 6.8 MPa from the above equation (4).

At this time, the pressure ratio P _s2 / P _{s1 of} the low pressure side compression element 11 is 1.9, the pressure ratio P _d / P _{s2 of} the high pressure side compression element 12 is 1.5, and the pressure difference P _s2 −P _{s1 of the} low pressure side compression element 11 = 3.3 MPa, and the pressure difference P _d −P _{s2 of} the high pressure side compression element 12 is 3.2 MPa.

  Therefore, the pressure ratio and pressure difference of each compression element approach, the load of each compression element and the leakage of the working chamber are reduced, and the efficiency of the compressor is improved.

On the other hand, when operating in the summer, the pressure conditions are as follows: suction pressure (low-pressure side suction pressure) P _s1 = 5.6 MPa, discharge pressure (high-pressure side discharge pressure) P _d = 10 MPa. The pressure ratio P _d / P _s1 = 1.8.

At this time, the second expansion valve 31 is controlled so that the discharge temperature difference (T _d2 −T _d1 ) between the high pressure side and the low pressure side becomes the target value shown in FIG. Then, the first expansion valve outlet pressure becomes lower than the outlet pressure of the low pressure side compression element 11, and about 25% of the refrigerant flows through the second refrigerant passage of the heat exchanger 20.

Therefore, the refrigerant flow ratio G ₂ / G ₁ = 0.75. Also, assuming a volumetric efficiency ratio of eta _v1 / eta _v2 = 1.0, and using a cylinder volume ratio V _th2 / V _th1 = 0.6 and an adiabatic index k = 1.3, the high-pressure side suction pressure is P _s2 = 7.5MPa.

At this time, the pressure ratio P _s2 / P _s1 = 1.3 of the low pressure side compression element 11, the pressure ratio P _d / P _s2 = 1.3 of the high pressure side compression element 12, and the pressure difference P _s2 −P _{s1 of the} low pressure side compression element 11 = 1.9 MPa, the pressure difference P _d −P _{s2 of} the high pressure side compression element 12 is 2.5 MPa.

  Therefore, the pressure ratio and pressure difference of each compression element approach, the load of each compression element and the leakage of the working chamber are reduced, and the efficiency of the compressor is improved.

  As in the present embodiment, the heat exchanger 20 has a structure in which the plurality of first refrigerant flow paths 21 and the second refrigerant flow paths 22 are in contact with the water passage 23 so that heat exchange is possible. Non-contact length can be shortened and heat exchange performance can be improved. Furthermore, if the water flow path 23 is comprised by an elliptical pipe | tube, the contact area of water and a pipe | tube can be enlarged, heat-transfer performance improves, and the efficiency of a heat pump type hot water heater can be improved. On the other hand, in refrigerants such as carbon dioxide where the heat exchanger has a supercritical state and the temperature decreases due to heat dissipation, the temperature difference between the water and the refrigerant is averaged by reversing the flow direction of the water and the refrigerant. Performance can be improved.

  The refrigeration cycle apparatus in the present embodiment includes a two-stage compressor having a low-pressure side compression element and a high-pressure side compression element, a heat exchanger that heats a cooling medium, a decompression device, and an evaporator, and the high-pressure side is super A first pressure reducing device and a second pressure reducing device connected in series as pressure reducing devices, and a second discharge passage connecting the outlet of the low pressure side compression element and the first pressure reducing device and the second pressure reducing device; Thus, the inlet pressure of the high pressure side compression element can always be properly controlled, and a highly efficient refrigeration cycle apparatus, for example, a heat pump apparatus or a refrigeration air conditioner can be provided.

  In the configuration of FIG. 10, the present invention is applied to a refrigerating and air-conditioning apparatus in which the heat exchanger 20 uses air as a cooling medium. In addition, a capillary tube 31b is provided as a second decompression device. 10, parts that are the same as those in FIG. 1 are given the same reference numerals, and explanation thereof is omitted.

  The heat exchanger 20 uses air as a cooling medium. More specifically, as shown in FIG. 11, the first refrigerant flow path 21 connected to the high pressure side compression element 12 via the suction passage 12a has a first refrigerant flow path inlet 21a downstream of the air flow, And it has the 1st refrigerant | coolant flow path exit part 21b connected with the 1st expansion valve 30 which is a 1st pressure reduction apparatus upstream of an air flow.

  The second refrigerant flow path 22 includes a second refrigerant flow path inlet 22a connected to the second discharge passage 11b 'through which the compressed gas from the low pressure side compression element 11 is guided, the first expansion valve 30, and the second pressure reducing device. It has the 2nd refrigerant | coolant flow path exit part 22b connected with piping between a certain capillary tube 31b. The heat exchanger 20 also includes fins 20a for promoting heat transfer.

The resistance of the capillary tube 31b is set as follows. That is, when it is operated at a low pressure ratio, such as summer conditions, and the suction gas temperature T _s1 detected by the temperature sensor 41, the degree of superheat is the difference between the evaporation temperature T _e detected by the temperature sensor 40 When the first expansion valve 30 is controlled so as to have a predetermined temperature, for example, 5 ° C., the pressure at the low pressure side compressor outlet is set to a predetermined value.

For example, the suction pressure P _s1 = 5.6 MPa, the discharge pressure (high-pressure side discharge pressure) P _d = 10 MPa, and the high-pressure side suction pressure is set to a resistance at which P _s2 = 7.5 MPa.

  With the above configuration, under the operating conditions with a large pressure ratio such as in winter, almost no refrigerant flows into the second refrigerant flow path 22 and is discharged from the low pressure side compression element 11 as in the configuration shown in FIG. The refrigerant thus drawn is sucked into the high pressure side compression element 12.

  Then, the refrigerant is compressed by the high-pressure side compression element 12 to become a higher pressure, and is branched at the first refrigerant flow path inlet 21 a and enters the heat exchanger 20.

  The high-pressure refrigerant flows through the first refrigerant flow path 21, is cooled by heating the air in the heat exchanger 20, and flows out of the heat exchanger 20 as a single refrigerant flow again at the first refrigerant flow path outlet 21b. To do.

  The refrigerant heat-exchanged in the first refrigerant flow path 21 of the heat exchanger 20 is depressurized by the first expansion valve 30 and the capillary tube 31 and then evaporated by the evaporator 32 and returns to the compressor 13.

  Under such a cycle condition, the refrigerant does not flow through the second refrigerant flow path 22, but the first refrigerant flow path 21 and the second refrigerant flow path 22 share the fins 20a. The fins 20b in the vicinity of the flow path 22 also work as heat transfer enhancement in the first refrigerant flow path 21, and there is little decrease in efficiency.

  Under the operation conditions with a small pressure ratio in summer as already described, the same operation as the configuration of FIG. 7 is performed. That is, since the resistance of the capillary tube 31 b is set so that the inlet pressure of the high-pressure side compression element 12 becomes appropriate, a part of the refrigerant that has come out of the low-pressure side compression element 11 passes through the second refrigerant flow path 22.

  The refrigerant gas that has flowed into the second refrigerant flow path 22 dissipates heat to the air, and merges with the refrigerant that has flowed out of the first refrigerant flow path 21 and has reached a low temperature, at the outlet of the first expansion valve 30. Then, the refrigerant gas returns to the compressor 13 through the capillary tube 31b and the evaporator 40.

  As described above, in the configuration of FIG. 10, the second refrigerant flow path 22 is connected between the first expansion valve 30 and the capillary tube 31b, so that the outlet of the high pressure side compression element is always at the outlet of the first expansion valve 30. The pressure can be made lower than the pressure, and the high-pressure side compression element works effectively to improve the efficiency.

  When air is used as the cooling medium, the inlet portions of the first refrigerant passage 21 and the second refrigerant passage 22 are leeward, and the outlet portions of the first refrigerant passage 21 and the second refrigerant passage 22 are By providing on the windward side, the temperature of the gas refrigerant in the supercritical state decreases from the leeward to the windward, and the temperature of the air increases from the leeward to the leeward, so that the heat exchange efficiency is improved.

Furthermore, in this embodiment, there is only one expansion valve, and control can be facilitated. Also the suction gas temperature T _s1 detected by the temperature sensor 41 in the present embodiment, control is performed in the superheat is a temperature difference between the evaporation temperature T _e detected by the temperature sensor 40, the high pressure side compression element You may control by the temperature of an exit. In this case, only one temperature sensor is required.

It is a refrigerating cycle block diagram which shows the outline of the heat pump type water heater which concerns on one Embodiment of this invention. It is a figure which shows the relationship of the temperature difference of the target discharge temperature difference, the discharge temperature of a high pressure side compression element, and the suction temperature of a low pressure side compression element. It is explanatory drawing of a refrigerant | coolant flow when the two-stage compressor of the heat pump type water heater of FIG. 1 performs high pressure ratio operation. It is explanatory drawing of a refrigerant | coolant flow when the two-stage compressor of the heat pump type water heater of FIG. 1 performs a low pressure ratio operation. It is a refrigerating cycle block diagram which shows the outline of the heat pump type water heater which concerns on other embodiment of this invention. It is a refrigerating cycle block diagram which shows the outline of the heat pump type water heater which concerns on other embodiment of this invention. It is a refrigerating cycle block diagram which shows the outline of the heat pump type heater based on one Embodiment of this invention. It is a refrigerating cycle block diagram which shows the outline of the heat pump type water heater which concerns on other embodiment of this invention. It is sectional drawing of the heat exchanger shown in FIG. It is a refrigerating cycle block diagram which shows the outline of the refrigerating air conditioning apparatus which concerns on further another embodiment of this invention. It is a block diagram of the heat exchanger shown in FIG.

Explanation of symbols

10 ... Compressor,
11 ... Low pressure side compression element,
11a: suction passage of the low pressure side compression element,
11b, 11b '... discharge passage of the low pressure side compression element,
12 ... high pressure side compression element,
12a ... suction passage of the high pressure side compression element,
12b: Discharge passage 13 of the high pressure side compression element ... Sealed container,
20, 20 '... heat exchanger,
21 ... 1st refrigerant path 22, 24a, 24b ... 2nd refrigerant path,
23 ... Water passage,
30 ... first expansion valve,
31 ... second expansion valve,
32 ... Evaporator,
33, 63 ... blower,
40 ... Evaporation temperature sensor,
41 ... Low-pressure side suction temperature sensor,
42 ... Low pressure side discharge temperature sensor,
43 ... High-pressure side discharge temperature sensor,
44 ... Pressure sensor in sealed container,
50 ... 1st controller,
51. Second controller,
60, 61, 62 ... cooler.

Claims (8)

  A compressor provided with an electric element housed in a hermetic container, a low-pressure side compression element and a high-pressure side compression element driven by the electric element, and heat exchange of the refrigerant discharged from the high-pressure side compression element A first heat exchanger, a decompression device that lowers the pressure of the refrigerant from the first heat exchanger, and a second heat exchanger that performs heat exchange of the decompressed refrigerant via the refrigerant flow path, respectively. A refrigerant circuit that sucks refrigerant from the second heat exchanger into the low-pressure side compression element of the compressor, and a refrigerant amount that is compressed by the low-pressure side compression element and sucked into the high-pressure side compression element. A refrigeration cycle apparatus comprising: a refrigerant amount adjusting unit for adjusting.   A compressor provided with an electric element housed in a hermetic container, a low-pressure side compression element and a high-pressure side compression element driven by the electric element, and heat exchange of the refrigerant discharged from the high-pressure side compression element The first heat exchanger, the first pressure reducing device for reducing the pressure of the refrigerant from the first heat exchanger, and the second heat exchanger for exchanging heat of the reduced pressure refrigerant respectively A refrigerant circuit that sucks the refrigerant from the second heat exchanger into the low-pressure side compression element of the compressor, the refrigerant compressed by the low-pressure side compression element, and the first decompressor An intermediate refrigerant passage communicating with the refrigerant passage between the second heat exchanger, a second decompression device communicating with the refrigerant from the intermediate refrigerant passage, and the first or / and the second decompression Control unit that controls the amount of refrigerant sucked into the high-pressure side compression element by controlling the device , The refrigeration cycle device having a.   3. The refrigeration cycle apparatus according to claim 2, wherein the second decompression device is provided on the intermediate refrigerant flow path.   In Claim 2, The said 2nd pressure reduction apparatus is provided on the refrigerant | coolant flow path connected with the said intermediate | middle refrigerant flow path, Comprising: The downstream of the flow of a refrigerant | coolant rather than the connection part with the said intermediate | middle refrigerant | coolant flow path. Refrigeration cycle equipment.   3. The refrigeration cycle apparatus according to claim 2, wherein the second pressure reducing device operates in accordance with a suction pressure of the high pressure side compression element or a discharge pressure of the low pressure side compression element.   3. The refrigeration cycle apparatus according to claim 2, wherein the second decompression device operates according to a difference between a discharge temperature of the high-pressure side compression element and a discharge temperature of the low-pressure side compression element.   3. The refrigeration cycle apparatus according to claim 2, wherein the intermediate refrigerant flow path includes a third heat exchanger that exchanges heat between the refrigerant compressed by the low-pressure side compression element. 8. The refrigeration cycle apparatus according to claim 7, wherein the first heat exchanger and the third heat exchanger exchange heat between the refrigerant flowing through each common fluid.

Images