GB2237625A

GB2237625A - Heat pump system

Info

Publication number: GB2237625A
Application number: GB9016544A
Authority: GB
Inventors: Shuhei Miyauchi; Toshimasa Irie; Tohru Isoda; Taizo Imoto; Yukio Fujishima; Yasuhiro Hatano; Masami Ogata; Yukitoshi Urata; Tamotsu Ishikawa; Masayuki Kawabata
Original assignee: NISHIYODO AIR CONDITIONER; Osaka Prefecture; Nishiyodo Air Conditioner Co Ltd
Current assignee: NISHIYODO AIR CONDITIONER; Osaka Prefecture; Nishiyodo Air Conditioner Co Ltd
Priority date: 1989-11-02
Filing date: 1990-07-27
Publication date: 1991-05-08
Anticipated expiration: 2010-07-27
Also published as: GB9016544D0; KR940009227B1; GB2237625B; JPH03148564A; CA2022125A1; DE4026699A1; AU610459B1; JP2552555B2; FR2653863B1; FR2653863A1; KR910010139A

Abstract

A heat pump system comprising a heat pump cycle of refrigerant which includes a compressor 1, a condenser 2, an expansion valve 3 and an evaporator 4, the condenser 2 further including a once-through path 6, 7 of fluid to be heated (e.g. water) and being a counterflow-type heat exchanger, the condenser being so operated that the refrigerant is supercooled in a supercooling degree of not less than 20% of the difference between a saturation temperature of the refrigerant and an inlet temperature of the fluid. The heat pump cycle may be a single cycle or a binary cycle (Fig. 1a) comprising a high- temperature size (1, 2, 3, 5) heat pump cycle of a high boiling refrigerant and a low-temperature side (11, 13, 14, 15) heat pump cycle of a low boiling refrigerant. The evaporator may include a once-through path for fluid to be cooled (e.g. water). Hot fluid and/or cold fluid are thus discharged. In particular, high quality hot water higher than an inlet temperature of water by 30-100 DEG C can be obtained. <IMAGE>

Description

HEAT PUMP SYSTEM

BACKGROUND OF THE INVENTION:

1. Field of the Invention:

This invention relates to a heat pump system including such a condenser that permits to discharge a high-quality fluid, i.e. extremely hightemperature fluid (e.g. water) with a high temperature difference between an inlet and an outlet temperatures of the fluid. More particularly, this invention provides a heat pump system which is characterized by heightening the supercooling degree of a refrigerant condensate in the condenser thereby to permit to increase its coefficient of performance and exchange heat quantity. 2. Statement of Prior Art:

Various heat pumps are known which are intended to discharge a hot fluid or cold fluid or both hot and cold fluids.

In general, a binary heat pump comprises, as shown in Fig..1a, a hightemperature side heat pump cycle for a high boiling refrigerant and a lowtemperature side heat pump cycle for a low boiling refrigerant whereby a hightemperature thermal output is discharged from the former cycle and a low-temperature thermal output, from the latter cycle. Examples of such binary heat pumps are disclosed in Japanese Patent First Publication Nos. 62-52376 (1987) and 62-52377(1987), Japanese Utility Model First Publication Nos. 57-2364 (1982) and 63-2053 (1988), etc. A heat pump includinga single heat pump cycle as shown in Fig. 1b is also A 2 - is known.

In such heat pump units, the coefficient of performance and attainable temperature of fluid to be heated usually vary, for a given refrigerant, depending upon the following factors: (a) performances of equipments such as a compressor, condenser, evaporator, etc., (b) conditions of two fluids conducting mutually heat exchange upon condensation, namely, the condensation tempe- rature of a refrigerant, and the condenser inlet temperature and flow rate of a fluid to be heated (e.g. water), (c.) evaporation conditions (flow rates and temperatures of two fluids effecting mutually heat exchange on the evaporation side), etc.

Hence the coefficient of performance of a heat pump and the attainable temperature of fluid to be heated are determined by the heat exchange conditions of the condenser side,.assuming that the refrigerant(s), the compressor(s), and the evaporator are predetermined and the evaporation conditions (flow rates and temperatures) are definite.

A conventional binary heat pump is operated in accordance with a refrigeration cycle presenting a temperature gradient chart as shown in Fig. 5a, wherein a refrigerant on the high-temperature side maintained in superheated gaseous state a at the inlet of the condenser (the outlet side of the compressor) is subjected to temperature change to reach b (saturated vapor line), is changed from the 3 - saturation vapor state b to saturation liquid state c within the condenser (b to c), and, at the outlet of the condenser (the inlet of an expansion valve), is changed to supercooled liquid state d (c to d).

On the other hand, its Mollier chart shows in Fig. 5b that the refrigerant is pressurized and compressed in the cOmpressor (f to a) with the enthalpy change of i'6 to ill; the resulting refrigerant vapor at superheated state passes over the saturation vapor line b with enthalpy change of if 1 to if 2' and is cooled by heat exchange with the fluid to be heated in the condenser and liquefied under a constant pressure (b to c), during which process the enthalpy changes from if 2 to "3; after passing over the point c (if 3 of the saturation liquid line, the refrigerant liquid becomes supercooled (c to d) resulting in the state of if 4' In Fig. 5a and Fig. 5b, the symbols t' wl and t' w2 designate a condenser inlet temperature and a condenser outlet temperature of the fluid to be heated, and the symbols e and f, an inlet state and oulet state of the cascade condenser.

In this manner, conventional refrigeration cycles have heretofore been operated so as to ensure a certain supercooling degree in order to operate the expansion valve without impairment, and the supercooling degree required to operate normally the expansion valve is, at the present day, considered to be on the order of 3 to 50C.

In this respect, with conventional heat pumps their 4 - i condensers have had a maximum heat transfer coefficient in the saturation zone of the refrigerant, but not so high heat transfer characteristics in the superheating zone and supercooling zone bf the refrigerant, and consequently, the condensers have been large-sized, which has led to decrease the economic merit. Further, with a view toward avoiding the reduction of coefficient of performance owing to the increase of pressure loss caused because of the large-sized condenser, a measure of making the supercooling degree great has not been considered to be advisable.

In the conventional heat pumps, heat exchangers of a non-counterflow type, such as shell-and-tube heat exchangers, heat exchangers of parallel flow type, etc. have been mostly adopted. A problem with these heat exchangers is that a lowest inlet temperature of water to be heated cannot be obtained, so that refrigerant liquid cannot sufficiently be cooled. For example, heat exchangers of parallel flow type, providing that the outlet water temperature is 90C, cannot lower the temperature of refrigerant liquid to less than 90 C. Even with a heat exchanger of a counterflow type, if a procedure of elevating gradually the temperature of water to be heated while circulating the water is taken, the inlet temperature of water for cooling the refrigerant liquid rises with time and is elevated to high temperature. For instance, if the inlet temperature of water is 90C in that case, the refrigerant liquid cannot be made lower than 9CC. When the flow rate of circulating water is made larger with the exchange heat quantity remained the same, the temperature - difference of water between the outlet and inlet temperatures cannot but be smaller. Consequently, there is still a problem that the inlet temperature is so high for the same outlet temperature that the refrigerant liquid cannot be cooled sufficiently.

In view of the various foregoing problems conventional heat pumps have posed, the present invention is designed to solve the problems particularly by taking account of the design of the condenser and its performance conditions. Accordingly, it is a principal object of this invention to provide a heat pump system presenting an increased coefficient of performance. Another object of this invention is to provide a heat pump system which permits to make a high-quality energy, viz. high-temperature steam or hot water available.

This invention suited to the foregoing objects is based on satisfying the following requirements for aquiring a high outlet temperature of water and securing a large supercooling degree: (a).adopting a heat exchange system of complete counterflow, (b) making a condenser inlet temperature of fluid to be heated (water) lower than a desired outlet temperature of refrigerant liquid from the condenser, (c) making the flow rate of the fluid to be heated relatively smaller as long as that's not inconvenient upon discharging.

SUMMARY OF THE INVENTION:

1 20 That is, according to one aspect of this invention, there is provided a heat pump system which includes a heat pump cycle for a refrigerant comprising a compressor, a condenser, an expansion valve and an evaporator, the condenser including a once-through path for a fluid to be heated, the condenser being constructed of a heat exchanger of thorough counterflow iype between the refrigerant and the fluid, the condenser being operated so that when the system is run, while delivering heat to the fluid in couterflow manner, the refrigerant is liquefied and condensed, and then supercooled in a supercooling degree of not less than 20% of the temperature difference between a saturation temperature of the refrigerant and a condenser inlet temperature of the fluid, whereby a hot fluid can be discharged from an outlet of the condenser.

In the aforesaid heat pump system, where the evaporator further includes a.once-through path for a fluid to be cooled and is of a heat exchanger of thorough counterflow type between the refrigerant and the fluid to be cooled, cold fluid only can be discharged or both hot and cold.fluids can be discharged.

According to another aspect of this invention, a binary heat pump system is provided which comprises a hightemperature side heat pump cycle for a high boiling refrigerant including a compressor, a condenser, an expansion valve and an accumulator connected by piping in the order mentioned, the condenser further including a once- 7 - through-path for a fluid to be heated, a low-temperature side heat pump cycle for a low boiling refrigerant including a compressor, an expansion valve, an evaporator and an accumulator connected by piping in the order mentioned, and a cascade condenser interconnecting the condenser of the high-temperature side cycle and the evaporator of the lowtemperature side cycle in a heat exchangeable manner, the condenser being of a heat exchanger of thorough counterflow type between the high boiling refrigerant and the fluid to be heated, the condenser being operated so that when the system is run, while conducting heat exchange with the fluid to be heated in counterflow manner, the high boiling refrigerant is liquefied and condensed and then supercooled in a supercooling degree of not less than 20% of the temperature difference between a saturation temperature of the refrigerant and a condenser inlet temperature of the fluid to be heated. whereby high temperature fluid can be discharged.

In the binary heat pump system above, where the evaporator further includes a once-through path for a fluid to be cooled, both hot fluid and cold fluid can be discharged.

In either case of the aforesaid heat pump units, the condenser is preferably formed of a double-tube or multitubular heat exchanger including an outer tube and an inner corrugated tube having wire fins, in which fluid to be heated is routed through the inner tube and the refrigerant is f 20 0 routed through the interspace between the inner and outer tubes in a counterflow manner to the fluid to be heated.

Throughout the specification and the claims, the term asupercooling degree" is intended to signify a temperature difference between a refrigerant saturation temperature and a condenser outlet temperature of the refrigerant liquid.

According to this invention, when the heat pump system is run, a refrigerant is evaporated in the evaporator, sucked into the compressor and discharged from there in gaseous state of high temperature and high pressure, liquefied as liquid condensate in the condenser whilegiving heat to a fluid to be heated flowing in counterflow to the refrigerant, then passes through the expansion valve to become gas-liqaid mixture state, and reverts to the evaporator. Here, in the condenser, for instance where the refrigerant-is s-dichlorotetrafluoroethane of high boiling point, it is supercooled in a supercooling degree of not less than 18.60C since its saturation temperature is ca. 112,C, and assuming that the inlet temperature of fluid to be heated is normal temperature, ca. 19.10C, the supercooling degree is calculated by: (112-19.1) x 0.2 = 18.6 (C). This value is significantly higher than the conventional supercooling degree of 3 to 5C. In general, the higher the supercooling degree, the higher is the coefficient of performance, and consequently, this invention apparently permits to enhance the coefficient of performance.

Furthermore, since the enthalpy difference of the t refrigerant between its saturation state in the condenser and its supercooling state at the outlet of the condenser is higher than that for conventional heat pumps,the temperature difference between the inlet and outlet temperatures of the fluid to be heated is higher than that of conventional heat pumps. As a consequence, high-temperature steam (on the order of 120OC) or hot fluid (ca. 10OcC) can be discharged with high efficiency. It is also possible to make hot fluid having a higher temperature than the saturation temperature of the refrigerant in the condenser available.

The refrigerant(s) which can be used for the heat pumps of this invention include, for example, fluorohydrocarbons such as 1,1,2-trichloro-1,2,2trifluoroethane (flon R-113), s-dichlorotetrafluoroethane (flon R-114), trichlorofluoromethane (flon R-11), dichlorodifluoromethane (flon R-12), chlorodifluoromethane (flon R-22), etc.

With a binary heat pump system, a high boiling refrigerant such as flon R-113, R-114, R-11, etc. can be used for the high-temperature side cycle whereas a low boiling refrigerant such as flon R-12, R-22, etc. can be used for the low-temperature side cycle.

BRIEF DESCRIPTION OF THE DRAWINGS:

Fig. la and Fig. 1b are schematic circuit diagrams of a fundamental binary heat pump unit and a single heat pump unit, respectively, to which this invefition is applicable.

Fig. 2a, Fig. 2b and Fig. 2c are a plan view, a side elevational view and a fragmentary enlarged view, respective- - 10 ly, of one example of a condenser for use in the heat pump system of this invention.

Fig. 3a and Fig. 3b are performance charts of binary heat pump systempertaining to this invention, Fig. 3a being a temperature gradient chart of one example of refrigerant and Fig. 3b,a Mollier chart thereof.

Fig. 4a. and Fig. 4b-are performance charts of another example of heat pump system pertaining to this invention, Fig. 4a being a temperature gradient chart of a heat pump cycle through another example of refrigerant and Fig. 4b, a Mollier chart thereof.

Fig. 5a and Fig. 5b are performance charts of a conventional heat pump operation based on the binary heat pump unit of Fig. la, Fig. 5a being a temperature gradient chart and-Fig. 5b, a Mollier chart.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS:

The invention.will be hereinbelow described in more detail with reference to the accompanying drawings.

This invention is applicable to a binary heat pump unit or single heat pump unit, depending upon the intended fluid to be discharged.

A binary heat pump unit shown in Fig. la usually comprises a high-temperature side cycle for high boiling refrigerant including a compressor 1, a condenser 2 having a once-through path for fluid to be heated, an expansion valve3 and an accumulator 5 connected in the order mentioned; and a low-temperature side cycle for low boiling refrigerant c i k, including a compressor 11, an expansion valve 13, an evaporator 14 having a once-through path for a fluid to be cooled, and an accumulator 15 connected in the order mention ed, the condenser 2 of the high-temperature side cycle and the evaporator 14 of the low-temperature side cycle being connected in a thermally exchangeable manner through a cascade condenser 22.

On the other hand, a single heat pump unit usually comprises, as shown in Fig. lb, a compressor 1, a condenser 2 having a once-through path for fluid to be heated, an expansion valve 3, an evaporator 4 having a once-through path for fluid to be cooled and an accumulator 5 connected in the order mentioned.

In either case, it is necessary that the fluid to be heated introduced into an inlet 6 of the condenser 2 from a fluid source be routed through the condenser in thorough counterflow to the.refrigerant while conducting heat exchange with it and be discharged from an outlet 7 of the condenser 2 in hot fluid or steam state.

It is preferred that the fluid to be cooled supplied from an inlet or._16 of the evaporator 4 or 14 and the refrigerant flow through the evaporator 4 or 14 in counter current manner. In the case of binary heat pump unit, it is preferred that the high boiling refrigerant and low boiling refrigerant flow through the cascade condenser 22 in counter current manner.

Fig. 3a and Fig. 3b show a temperature gradient chart - 12 and a Mollier chart, respectively, of a binary heat pump system according to this invention, wherein the high boiling refrigerant is sdichlorotetrafluoroethane (flon R-114), and are comparable with Fig. 5a and Fig. 5b of conventional-heat pump operation. As will be apparent from the comparison of the tempera- ture grdient (Fig. 3a and Fig. 5a), in accordance with the invention, the high boiling refrigerant maintained in super heated gaseous state (A) at the inlet of the condenser 2 is changed inside the condenser from its saturated gaseous state to saturated liquid state (B to C), which change of state is analogous to that of conventional operation method shown in Fig. 5a. However, significant changes can be seen in that the refrigerant in saturated liquid state is changed at the outlet of the condenser 2 to supercooling state (D), exhibiting a larger supercooling degree and that the tempe rature difference on the water side between the outlet (tw2) and inlet (t wl) of the condenser is larger as compared with conventional operation.

A heat pump cycle according to this invention is represented in a Mollier chart of Fig. 3b wherein in the course during which the refrigerant in superheating state (A) at the inlet 6 of the condenser 2 becomes saturated gaseous state (B), the enthalpy is changed from il to '2; in the course during which the refrigerant is further cooled by heat exchange with water (fluid to be heated) and liquefied and condensed at a constant pressure inside the condenser 2 to be saturated liquid (B to C), the enthalpy is changed to i 3 and when the refrigerant is in supercooled state (D) at the outlet 7 of the condenser 2, an enthalpy of i 4 is reached. The enthalpy change of from i 3 to i 4 is thus much increased in Fig. 3b of this invention, as compared with that of Fig. 5b, which fact shows an increase in supercooling degree. Then, in the throttling expansion course of D to E, the refrigerant flows through the expansion valve 3 into the cascade condenser 22 in the same enthalpy of i 4 = i 5 and there, the refrigerant is evaporated completely (E to F) to become the state of i 6 The refrigerant having an enthalpy of i 6 is sucked into the compressor 1.

0 20 Figs. 2a to 2c illustrate details of one example of-a condenser used for the present heat pump system, which condenser is formed of a double tube 30 comprising a corrugated inner tube 32 having wire fins 33 and an outer tube 33. A refrigerant can flow through the interstices between the inner tube 32 and the outer tube 33 from the upper side whereas fluid to be heated can flow through the inner tube 32 from the lower side.

In this manner, according to the heat pump system of this invention, a high-quality hot water or steam having an extremely high temperature on the order of 120C can be discharged with a large temperature difference of 30 to 100 OC, usually 50 to 900C between the condenser inlet and outlet temperatures. A cold water on the order of 7C can be made available solely or simultaneously with hot water.

14 p Several examples of this invention will be shown below. Example 1 Binary heat pump unit as illustrated in Fig. la was operated by the use of a condenser having a construction shown in Table 1 below, water as fluid to be heated and fluid to be cooled, and flon R-114 and R-22 as refrigerants for high-temperature side cycle and low-temperature side cycle, respectively, under the conditions shown in Table 2 below. Further physical data are also shown in Table 2. Table 1 Heat Transfer Tube Shape Outer Tube (Diameter) Inner Tube (Diameter) Length Heat Transfer Area Corrugation Pitch Corrugation Depth Height of Wire Fins Pitch of Wire Fins Wire Fin Corrugated Tube Double Tube 25.4 OD X t 1.2 x 23.0 ID Mm 12.7 OD X t 1.7 x 11.3 ID mm 3 634 m 0.154 m2 4.67 mm 0.21 mm 0.8 mm 0.48 mm j j, Table 2

Condenser Superheating Zone Saturation Zone Supercooling Zone 0 Exchange Heat Quantity(kcal/h) Condenser Inlet Temp. of Water ( 11 C) Condenser Outlet Temp. of Water ( "C) Condenser Outlet Temp. of Refrigerant (-C) Saturation Temp. of Refrigerant C,C) Superheating Degree VC) Supercooling Degree CC) Flow Rate of Water (liter/h) Flow Rate of Refrigerant (kg/h) Quantity of Heat (kcal/h) Overall Heat Transfer Coefficient (kcal/m2h0C) Heat Transfer Coefficient on the Refrigerant Side (kcallm2hC) Heat Transfer Coefficient on the Water Side (kcal/M2hC) Heat Transfer Area Percentage(%) 9552 19.1 98.7 59.5 112 7.1 52.5 275.3 496 1131 1449 5671 5122 3260 10859 5124 17.4 34.1 3937 1246 1929 3873 48.5 Notes: Exchange Heat Quantity = Flow Rate of Water x (Outlet Temp. of Water - Inlet Temp.'of Water Supercooling Degree = Saturation Temp. of Refrigerant Outlet Temp. of Refrigerant - 16 Properties of the high boiling refrigerant (flon R-114) in the heat pump cycle according to this invention were measured, and values of Mollier diagram obtained are shown in-Table 3 below, in comparison with those of conventional heat pump cycle.

Table 3

This Inventio;k-tate A B c D E F Temperature (C) 119.1 112 -112 59.5 35 78 Pressure (kgfjcm2) 18.2 18.2 18.2 18.2 3.0 3.0 Enthalpy (kcal/kg) i 1 '2 i 3 i 4 i 5 i 6 148.8 147.0 128.4 114.1 114.1 145.4 Conventional",,tate a b c d e f Temperature (OC) 119.1 112 112 107 35 78 Pressure (kgf/CM2 18.2 18.2 18.2 18.2 3.0 3.0 Enthalpy (kcallkg) it 1 if 2 if 3 if 4 if 5 if 6 148.8 147.0 128.4 127.1 127.1 145.4 Notes.:

i The symbols of "A" to "F" and "a" to "f" correspond to the Mollier diagrams of Fig. 3b and Fig. 5b, respectively.

From Table 3 above, the following values are calculated.

Supercooling Temperature cop3 Degreel Effectiveness2 This Invention 52.5"C Conventional 5C Notes:

1 Supercooling Degree = TC R 56.5% 5.4% - TD or T c - T d 10.2 6.4 - 17 1 8 2 Temperature Effec tiveness c D - or c d T c t W1 T c t, wl 3 Coefficient of Performance = il - i4 is 1 - i, 4 or - 1 1 - 1 6 i c 1 - i 0 6 From Table 3, it will be apparent that the enthalpy difference of the refrigerant liquid upon subcooling is greater in this invention than in the conventional heat pump operation.

Further, the relation between supercooling degree of the refrigerant (R114) in the condenser and coefficient of performance was examined, and the results obtained are shown in Table 4 below.

The measurement conditions are as follows:

Saturation Pressure 18.2 kgf/cm2 Saturation Temperature (T c) 112.0C Condenser Inlet Temperature of Water 19.1C (twl) Enthalpy at Compressor Inlet (i 6) 145.4 kcal/kg Enthalpy at Compressor Outlet(i 1) 148.8 kcallkg Table 4

Temperature Effectivenessl(%) 10 20 30 40 Supercooling Degree2(C) 4.6 9.3 18.6 27.9 37.2 Outlet Temp. of Refrigerant Liq.

T D ("C) 107.4 102.7 93.7 84.1 74.8 Enthalpy of Refrigerant Liq. at OutL i "kcallkg) 41 127.1 125.6 122.9 120.4 118.0 Coefficient of Performance3 6.4 6.8 7.6 8.4 9.1 18 - 60 70 80 48.4 65.6 56.3 47.0 55.7 65.0 74.3 37.7 115.7 113.4 ill. 1 108.9 9.7 10.4 11. 1 11.7 Notes 1 Temperature Effectivexiess c D T c t wl 2 Supercooling Degree = T c - T D 112 - T D 3 Coefficient of Performance 1 4 1 6 1 112 - T D 92.9 148.8 - i 3.4 At the outlet 7 of the condenser 2, hot water of ca. 99 OC was discharged with a temperature difference of ca. 80C whereas at the outlet 19 of the evaporator 14, cold water of 70C was discharged from the inlet temperature of 120C. Example 2 A heat pump unit as shown in Fig. lb was run by using dichlorodifluoromethane (R-12) as refrigerant, a condenser of the construction shown in Table 5 below and water as fluids to be heated and cooled, under the conditions in Table 6 below. The resulting data are also shown in Table 6. Table 5 Heat Transfer Tube Outer Tube (Diameter) Inner Tube (Diameter) Wire Fin Corrugated Tube (Double-tube) 31.8 OD X t 1.6 x 30.2 ID mm 19.05 OD X t 0.95 x 17.15 ID rm Length Heat Transfer Area Corrugation Pitch 3 520 m x 4 0.84 m 2 7. 2 mm 1 Corrugation Depth Height of Fins Fin Pitch 0.31 mm 0. 8 mm 0.48 mm Table 6

Condenser Super- Saturation Super heating Zone cooling Zone 1 Zone 13,630 20.4 Exchange Heat Quantity (Kcal/h) Condenser Inlet Temp. of Water (OC) Condenser Outlet Temp. of Water V,C) Saturation Temp. (OC) Superheating Degree (OC) Supercooling Degree (C) Flow Rate of Water (liter/h) Flow Rate of Refrigerant (kg/h) Quantity of Heat (kcal/h) Difference between Outlet Temp. and Inlet Temp. of Water (C) 0 96.2 84.6 50.6 46.6 303.9 6470 3370 18.7 36.0 3790 21.1 The temperature gradient and Mollier diagram of this heat pump cycle are diagramatically shown in Fig. 4a and Fig 4b, respectively.

Properties of R-12 in the heat pump cycle presenting the Mollier diagram (Fig. 4b) are shown in Table 7 in comparison with those of conventional heat pump cycle.

This Invention\State Temperature C Pressure kgfjcm2 Enthalpy kcallkg Conventional\state Temperature OC Pressure kgf /cm2 Enthalpy keal/kg Table 7

A B C D 135.2 84.6 84.6 38.0 25.6 25.6 25.6 25.6 i 1 i 2 i 3 i 4 153.8 142.7 121.4 108.9 a b C d e f 135.2 84.6 84.6 79.6 0.49 30.1 25.6 25.6 25.6 25.6 3.2 3.2 it 1 if 2 if 3 if 4 it 5 if 6 153.8 142.7 121.4 119.9 119.9 141.0 E F 0.49 30.1 3.2 3.2 i 5 i 6 108.9 141.0 Notes:

The symbols A to F designate the states of Fig. 4b whereas the symbols a to f designate corresponding states of Mollier chart to Fig. 5b in the conventional heat pump system.

From Table 7 above, the following values of performances are calculated.

1 Supercooling Temperature Degree Effectiveness COP This Invention 46.6"C 72.6% 3.5 Conventional 50C 7.8% 2.6 In this way, hot water of ca. WC was available with a temperature difference of ca. 760C and cold water of 70C was available with a temperature difference of 5C.

Thus far descr-ibed, this invention provides a heat pump system havingsuch a heat pump cycle that in the condenser, 3 J 1 - 21 the refrigerant is subjected to heat exchange in counterflow manner with fluid to be heated and the refrigerant condensate is supercooled in a supercooling degree of not less than 20% of the temperature difference between a refrigerant saturation temperature and a condenser inlet temperature of the fluid, whereby enthalpy difference of the refrigerant liquid upon supercooling is increased and temperature difference of the fluid between the inlet and outlet temperatures is increased. Accordingly, the exchange heat quantity, which is determined depending upon the flow rate of and difference between the inlet and outlet temperatures of the fluid, is increased even with a small flow rate of the fluid. Hence it is possible to much heighten the outlet temperature of the fluid such as water as well as the heat exchange efficiency.

Such high supercooling degree and an increased enthalpy difference in the. supercooling zone, attended by the counterflow process, permit to enhance the coefficient of performance and to make a high-quality hot Water having an extremely high temperature of 120C available instantaneously. Moreover, it is possible to lessen the capacity of the compressor because the flow rate of refrigerant can be decreased, thus curtailing the installation cost and running cost.

1

Claims

What is claimed is:

1. A heat pump system comprising a beat pump cycle of refrigerant including a compressor, a condenser, an expansion valve and an evaporLtor all connected in the order mentioned, said condenser further including a once-through path for a fluid to be heated, said condenser being constructed of a heat exchanger of thorough counterflow type between the refrigerant and fluid, said condenser being operatedwhen the system is run,so that while conducting heat exchange with the fluid in counterflow manner, the refrigerant is liquefied and condensed and subsequently supercooled in a supercooling degree of not less than 20% of the difference between a saturation temperature of the refrigerant and an inlet temperature of the fluid, whereby hot fluid can be discharged from an outlet of the condenser.

2. The heat pump system as set forth in claim 1, wherein said evapo.rator further includes a once-through path for a fluid to be cooled, whereby cold fluid or hot fluid and cold fluid can be discharged.

3. The heat pump system as set forth in claim 2, wherein said evaporator is constructed of a heat exchanger of thorough counterflow type between the refrigerant and the fluid to be cooled.

4. The heat pump system as set forth in claim 1, wherein said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid by 30 to 1OCC.

t t 1

5. The heat pump system as set forth in claim 1, wherein _said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid by 50 to OC.

6. The heat pump system as set forth in claim 1, wherein said heat exchanger is a double-tube or multi-tube heat exchanger.

7. The heat pump system as set forth in claim 1, wherein said condenser is operated so that ah outlet temperature of the fluid is higher than a saturation temperature of the refrigerant.

8. A heat pump system comprising a-high-temperature side heat pump cycle of high boiling refrigerant including a compressor, a condenser and an expansion valve; a lowtemperature side heat pump cycle of low boiling refrigerant including a compressor, an expansion valve and an evaporator; and a cascade condenser interconnecting both cycles located between the compressor and expansion valve of the former cycle and the compressor and expansion valve of the latter cycle; said condenser further including a once-through path for a fluid to be heated and being constructed of a heat exchanger of thorough counterflow type between the high boiling refrigerant and the fluid, said condenser being operated, when the system is.run, so that while conducting heat exchange with the fluid in counterflow manner, the high boiling refrigerant is liquefied and condensed and then - 24 supercooled in a supercooling degree of not less than 20% of the difference between a saturation temperature of the high boiling refrigerant and a condenser inlet temperature of the fluid, whereby hot fluid can be discharged from an outlet of the condenser.

9. The heat pump system as set forth in claim 8, wherein said evaporator further includes a once-through path for fluid to be cooled, whereby hot fluid and cold fluid can be discharged simultaneously.

10. The heat pump system as set forth in claim 9, wherein the evaporator is of a heat exchanger of thorough counterflow type between the low boiling refrigerant and the fluid to be cooled, and the cascade condenser is of a heat exchanger of thorough counterflow type between both refrigerants.

11. The heatpump system as set forth in claim 8, wherein said heat exchanger is a double-tube or multi-tube heat.exchanger.

12. The heat pump system as set forth in claim 8, wherein said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid by 30 to 1000C.

13. The heat pump system as set forth in claim 8, wherein said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid by 50 to 900C.

14. The heat pump system as set forth in claim 9, - 25 wherein said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid to be heated by 50 to WC.

15. A heat pump system substantially as hereinbefore described with reference to: Figure la or Figure lb; or Figure la or Figure 1b together with any of Figures 2a to 2c; Figures 3a and 3b; or Figures 4a and 4b of the accompanying drawings.

i Published 1991 at Tbe Patent Office, State House. 66/71 High Holborn. London WC I R47?. Further copies rnay be obtained frorn Sales Branch, Unit 6. Nine Mile PbInt, Cwmiclinfach. Cross Keys. Newport. NP I 7HZ. Printed by Multiplex techniques lid, St Mary Cray. Kent.