AU610459B1

AU610459B1 - Heat pump system

Info

Publication number: AU610459B1
Application number: AU59877/90A
Authority: AU
Inventors: Yukio Fujishima; Yasuhiro Hatano; Taizo Imoto; Toshimasa Irie; Tamotsu Ishikawa; Tohru Isoda; Masayuki Kawabata; Shuhei Miyauchi; Masami Ogata; Yukitoshi Urata
Original assignee: Osaka Prefecture; Nishiyodo Air Conditioner Co Ltd
Current assignee: Osaka Prefecture; Nishiyodo Air Conditioner Co Ltd
Priority date: 1989-11-02
Filing date: 1990-07-26
Publication date: 1991-05-16
Anticipated expiration: 2010-07-26
Also published as: KR910010139A; KR940009227B1; GB2237625A; GB2237625B; CA2022125A1; JP2552555B2; GB9016544D0; FR2653863A1; DE4026699A1; FR2653863B1; JPH03148564A

Description

GRIFFITH HACK CO PATENT AND TRADE MARK ATTORNEYS MEL O URN E SYDNEY PER TH

AUSTRALIA

PATENTS ACT 1952 COMPLETE SPECIFICATION Form

(ORIGINAL)

FOR OFFICE USE Short Title: Int. Cl: Application Number: Lodged:

I

Complete Specification-Lodged: Accepted: Lapsed: Published: Priority: Related Art: I This documint contains the iamnndmrents made under Section 49 and is correct for Sprinl ing. TO BE COMPLETED BY APPLICANT Name of Applicant: Address of Applicant: Address of Applicant: 1) OSAKA PREFECTURE 1-22, OHTEMAE 2-CHOME CHUO-KU, OSAKA

JAPAN

2) NISHIYODO AIR CONDITIONER CO., LTD.

15-10, HIMESATO 1-CHOME

NISHIYODOGAWA-KU

OSAKA

JAPAN

Actual Inventor: Address for Service: GRIFFITH HACK CO., 601 St. Kilda Road, Melbourne, Victoria 3004, Australia.

Complete Specification for the invention entitled: HEAT PUMP SYSTEM.

The following statement is a full description of this invention including the best method of performing it known to me:- 1 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 high-temperature fluid 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 tharacterized o 0 a o 0 o. by heightening the supercooling degree of a refrigerant o 0o O.o its coefficient of performance and exchange heat quantity.

oo 00 0 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 0 0 0"5 fluids.

000 o In general., a binary heat pump comprises, as shown in 0o, Fig. .la, a high-temperature side heat pump cycle for a high boiling refrigerant and a low-temperature side heat pump cycle for a low boiling refrigerant whereby a high- °3°020 temperature thermal output is discharged from the former 00 o 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 including a single heat pump cycle as shown in Fig. Ib is also c I I YT -2known.

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: performances of equipments such as a compressor, condenser, evaporator, etc., conditions of two fluids conducting mutually heat exchange upon condensation, namely, the condensation tempe- O ~rature of a refrigerant, and the condenser inlet temperature and flow rate of a fluid to be heated water), evaporation conditions (flow rates and temperatures of 000 Sa two fluids effecting mutually heat exchange on the 0 00 oe evaporation side;, etc.

Hence the coefficient of performance of a heat pump and 0 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.

o °A conventional binary heat pump is operated in o 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 UT-~ I I I I 3 saturation vapor state b to saturation liquid state c within the condenser (b to 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. that the refrigerant is pressurized and compressed in the compressor (f to a) with the enthalpy change of i! 6 to i'l; the resulting refrigerant vapor at superheated state passes over the saturation vapor line b with enthalpy change of i'1 to 0 o~10 i' 2, and is cooled by heat exchange with the fluid to be oa heated in the condenser and liquefied under a constant 0 Qa pressure (b to during which process the enthalpy changes (00 0 a 0 from i' 2 to i after passing over the point c (i' 3 of the saturation liquid line, the refrigerant liquid becomes supercooled (c to d) resulting in the state of i' 4 oIn Fig. 5a and Fig. 5b, the symbols t'wl and t'w2 designate a condenser inlet temperature and a condenser o a outlet temperature of the fluid to be heated, and the symbols 00 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 In this respect, with conventional heat pumps their

I

I 4condensers 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 of 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, ,I heat exchangers of parallel flow type, etc. have been mostly t: 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 1 cooled. For example, heat exchangers of parallel flow type, providing that the outlet water temperature is 90 0 C, cannot lower the temperature of refrigerant liquid to less than 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 90°C in that case, the refrigerant liquid cannot be made lower than When the flow rate of circulating water is made larger with the exchange heat quantity remained the same, the temperature rr-rrr~ 5 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.

According to the present invention there is provided a heat pump system comprising a heat pump cycle of refrigerant including a compressor, a condenser, an expansion valve and an evaporator 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 operated, when the system is run, so that while conducting heat exchange with the fluid in counterflow manner, the refrigerant is liquefied and o0 condensed and subsequently supercooled in a supercooling aw degree of not less than 20% of the difference between a o o saturation temperature of the refrigerant and an inlet temperature of the fluid, whereby hot fluid can be discharged from an outlet of the condenser.

o 0 According to the present invention there is further provided a heat pump system comprising a high-temperature side heat pump cycle for high boiling refrigerant including a compressor, a condenser and an expansion valve; a low-temperature side heat pump cycle for low boiling 000 refrigerant including a compressor, an expansion valve and 0 an evaporator; and a cascade condenser interconnecting both 94 cycles located between the compressor and expansion valve of the.h.igh-temperature side heat pump cycle and the 0O. compressor and expansion valve of the low-temperature side *4*4 heat pump cycle; 00 4 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 6 conducting heat exchange with the fluid 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 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.

BRIEF DESCRIPTION OF THE DRAWINGS In order that the present invention might be more fully understood, embodiments of the invention will be described by way of example only with reference to the accompanying drawings in which: Fig. la and Fig. lb are schematic circuit diagrams of a fundamental binary heat pump unit and a single heat 0a pump unit, respectively, in accordance with an embodiment of the invention.

~oo Fig. 2a, Fig. 2b and Fig. 2c are a plan view, a o 0o side elevational view and a fragmentary enlarged view, respectively, of one example of a condenser for use in the Oheat pump system of an embodiment of this invention.

0 0 o a o 1 ~Fig. 3a and Fig. 3b are performance charts of binary heat pump system pertaining to an embodiment of this invention, Fig. 3a being a temperature gradient chart of one example of refrigerant and Fig. 3b, a Mollier chart 000 1 thereof.

St o Fig. 4a and Fig. 4b are performance charts of oOO 1 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 0 0 0 Fig. 4b, a Mollier chart thereof.

.000 o i 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.

0 a 00 0 00O 00 0 00 00 0 0 00 ao o 0 a 000 0 0 0 0 00 Oao 0 0 01D o a 0 0I 0 0Q 0 Q O 0 S00 0 00 7 DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS Embodiments of the invention will be hereinbelow described in more detail with reference to the accompanying drawings.

Embodiments of this invention are applicable to a binary heat pump unit or single heat pump unit, depending upon the intended fluid to be discharged.

The following factors relate to the acquiring of a high outlet temperature of water and securing a large super-cooling degree: adopting a heat exchanger system of complete counterflow, making a condenser inlet temperature of fluid to be heated (water) lower than a desired outlet temperature of refrigerant liquid from the condenser, making the flow rate of the fluid to be heated relatively smaller as long as that's not inconvenient upon discharging.

A heat pump system, in accordance with an embodiment of the invention, 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 type between the refrigerant and the fluid, the condenser being operated so that when the system is run, while delivering heat to the fluid in counterflow manner, the refrigerant is liquefied and condensed, and then supercooled in a supercooling degree of not less than 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 FS c I-i.r~-n 8 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.

A binary heat pump system can be provided which comprises a high-temperature 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-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 00evaporator of the low-temperature side cycle in a heat *exchangeable manner, the condenser being of a heat 0o exchanger of thorough counterflow type between the high 0 000 0 boiling refrigerant and the fluid to be heated, the 0 0. 0 o00o condenser being operated so that when the system is run, o00°0 while conducting heat exchange with the fluid to be heated in counterflow manner, the high boiling refrigerant is 0 00 o 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 0 0 0-0 to be heated, whereby high temperature fluid can be 11 0 o discharged.

0 "o In the binary heat pump system above, where the 0 0 evaporator further includes a once-through path for a fluid to be cooled, both hot fluid and cold fluid may be 0o° discharged simultaneously.

t oe In either case of the aforesaid heat pump units, 0 the condenser is preferably formed of a double-tube or multi-tubular heat exchanger including an outer tube and an inter corrugated tube having wire fins, in which fluid to be heated is routed through the inner tube and the 1 iP~ -9refrigerant is routed through the interspace between the inner and outer tubes in a counterflow manner to the fluid to be heated.

The term "supercooling degree" is intended to signify a temperature difference between a refrigerant saturation temperature and a condenser outlet temperature of the refrigerant liquid.

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 while giving heat to a fluid to be heated flowing in counterflow to the refrigerant, then passes through the expansion valve to become gas-liquid mixture state, and reverts to the evaporator. Here, in the 0 condenser for instance where the refrigerant is o s-dichlorotetrafluoroethane of high boiling point, it is o 0 o supercooled in a supercooling degree of not less than 018.6 0 C since its saturation temperature is ca. 112°C, and 0 assuming that the inlet temperature of fluid to be heated o is normal temperature, ca. 19.1 0 C, the supercooling degree °0 is calculated by: (112-19.1) x 0.2 18.6 This value 0 6 is significantly higher than the conventional supercooling degree of 3 to 5 0 C. In general, the higher the supercooling degree, the higher is the coefficient of performance, and consequently, the coefficient of 0 performance can be enhanced.

Furthermore, since the enthalpy difference of the 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 4 4 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 1200C) or hot fluid (ca. 100 0 C) can be discharged with high efficiency. It is also possible to make hot fluid having a higher temperature 'g LLtQ y

A

-r I _19 1 10 than the saturation temperature cf tha refrigerant in the condenser available.

The refrigerant(s) which can be used for the heat pumps include, for example, fluorohydrocarbons such as 1,1,2-trichloro-1,2,2-trifluoroethane (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.

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 °o valve 3 and an accumulator 5 connected in the order o mentioned; and a low-temperature side cycle for low boiling o0 refrigerant o0o 0 00 0 0 o >0 0 O K o a 0 00a is normal temperature, ca. 19.1-C, the supercooling aegree is calculated by: (112-19.1) x 0.2 18.6 This value /2 4 I i i

I~IIIII~

I s 11

CI

0 0~ 0

C

0 0 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 mentioned, 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. Ib, a compressor i, a condenser 10 2 having a once-through path for fluid to be heated, an expansion valve 3, an evaporator 4 having a once-through S path for fluid to be cooled and an accumulator 5 connected in the order mentioned.

444 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 6 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 o from an inlet 6 or 16 of the evaporator 4 or 14 and the o refrigerant flow through the evaporator 4 or 14 in countercurrent 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 countercurrent manner.

Fig. 3a and Fig. 3b show a temperature gradient chart 0 0 Onl 4 0 0 saia conaenser rurtner including a once-through path for a fluid to be heated and being constructed of a /3 12 and a Mollier chart, respectively, of a binary heat pump system according to this invention, wherein the high boiling refrigerant is s-dichlorotetrafluoroethane (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 temperature grdient (Fig. 3a and Fig. 5a), in accordance with the invention, the high boiling refrigerant maintained in superheated gaseous state at the inlet of the condenser 2 is i 0 changed inside the condenser from its saturated gaseous V4 state to saturated liquid state (B to which change of state is analogous to that of conventional operation method 4 4 shown in Fig. 5a. However, significant changes can be seen 4 4 in that the refrigerant in saturated liquid state is changed at the outlet of the condenser 2 to supercooling state exhibiting a larger supercooling degree and that the tempe- 44. .rature difference on the water side between the outlet (tw 2 and inlet (twl) 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 K4 course during which the refrigerant in superheating state at the inlet 6 of the condenser 2 becomes saturated gaseous state the enthalpy is changed from ii to i 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

LI---_L

t' 'I I 13 to be saturated liquid (B to the enthalpy is changed to i3, and when the refrigerant is in supercooled state 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 1 0 there, the refrigerant is evaporated completely (E to F) to O become the state of i 6 The refrigerant having an enthalpy 04 of i 6 is sucked into the compressor 1.

0 Figs. 2a to 2c illustrate details of one example of a 0. 0 condenser used for the present heat pump system, which condenser is formed of a double tube 30 comprising a corrugat-_J 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 0 o. upper side whereas fluid to be heated can flow through the inner tube 32 from the lower side.

0*a In this manner, according to the heat pump system of o this invention, a high-quality hot water or steam having an extremely high temperature on the order of 120°C can be discharged with a large temperature difference of 30 to 100 usually 50 to 90°C between the condenser inlet and outlet temperatures. A cold water on the order of 7 0 C can be made available solely or simultaneously with hot water.

I c 1 I d 14 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 3 5 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.

0 Table 1 1 t

I'

a o I @4, Heat Transfer Tube Wire Fin Corrugated Tube i Shape Outer Tube (Diameter) Inner Tube (Diameter) Length Heat Transfer Area Corrugation Pi~ch Corrugation Depth Height of Wire Fins Pitch of Wire Fins Double Tube CD t ID 25.4 x 1.2 x 23.0 mm OD t ID 12.7 x 1.7 x 11.3 mm 3 634 m 0.154 m 2 4.67 mm 0.21 mm 0.8 mm 0.48 mm

I

I I I 4 44 44 44 4 3 6 a 0 4 404 o a 44 o44 444 64 a 4 o a 0 44 oa a a 04 4 4 4 4 44 44a 15 Table 2 Condenser Super- Saturation Superheating Zone cooling Zone Zone Exchange Heat Quantity*(kcal/h) 9552 Condenser Inlet Temp. of Water 19.1

(OC)

Condenser Outlet Temp. of 98.7 Water (OC) Condenser Outlet Temp. of 59.5 Refrigerant

(OC)

Saturation Temp. of Refrigerant 112 Superheating Degree 7.1 Supercooling Degree** 52.5 Flow Rate of Water (liter/h) 120 Flow Rate of Refrigerant (kg/h) 275.3 Quantity of Heat (kcal/h) 496 5122 3937 Overall Heat Transfer 11 Coefficient (kcal/m 2 h°C) 1131 3260 1 Heat Transfer Coefficient on the 1449 10859 1929 Refrigerant Side (kcal/m 2 h°C) Heat Transfer Coefficient on the Water Side (kcal/m 2 hC) 5671 5124 3873 Heat Transfer Area Percentage(%) 17.4 34.1 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

_L~

r liii 00 00 0 00 0 I 00 o 0 0 00 0a oo 00 0 00 000 0 0.

a a a t3 o 0 00 0 0 O 00 0 0 00« 00 .0 0 0 oa "ld 0 0 0 0 a 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 Invention tate A B C D E F Temperature (OC) 119.1 112 112 59.5 35 78 Pressure (kgf/cm 2 18.2 18.2 18.2 18.2 3.0 Enthalpy (kcal/kg) i i 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 119.1 112 112 107 35 78 Pressure (kgf/cm 2 18.2 18.2 18.2 18.2 3.0 Enthalpy (kcal/kg) i' 1 i' 2 i' 3 i' 4 i 5 i 6 148.8 147.0 128.4 127.1 127.1 145.4 Notes': The symbols of to and to correspond to the Mollier diagrams of Fig. 3b and Fig. 5b, respectively.

From Table 3 above, the following values are calculated.

Supercooling Tem Degree*l Effec This Invention 52.5"C 5 Conventional 5°C Notes: *1 Supercooling Degree TC TD perature COP*3 tiveness*2 6.5% 10.2 5.4% 6.4 or T T c d 17 T T T T *2 Temperature Effectiveness or T twl T t' cC wwl c Wl *3 Coefficient of Performance i4 or i 1 i4 1 6 1 4 or 1 6 i i' 1 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 0 oa S' the refrigerant (R-114) in the condenser and coefficient of 0o o performance was examined, and the results obtained are shown 0,4( in Table 4 below.

SThe measurement conditions are as follows: Saturation Pressure 18.2 kgf/cm2 Saturation Temperature (TC) 112.0°C Condenser Inlet Temperature of Water 19.1°C (tw l) Enthalpy at Compressor Inlet i 6 145.4 kcal/kg Enthalpy at Compressor Outlet(i 148.8 kcal/kg Table 4 0 00 000 0 0 6 0l I 0 P Temperature Supercooling Outlet Temp. Enthalpy of Coefficient Effective- Degree*2(°C) of Refrige- Refrigerant of Perforness*l(%) rant Liq. Liq. at Out- mance*3 TD il4ekcal/kg) 4.6 107.4 127.1 6.4 9.3 102.7 125.6 6.8 18.6 93.7 122.9 7.6 27.9 84.1 120.4 8.4 37.2 74.8 118.0 9.1 'i I i I. d 18 Notes 48.4 55.7 65.0 74.3 65.6 56.3 47.0 37.7 115.7 113.4 111.1 108.9 9.7 10.4 11.1 11.7 00 00 o0

O

00 0 04 a oa 0 S 0 6 0 0 0 i T T 112 T C D D *1 Temperature Effectiveness

T

C t1 92.9 *2 Supercooling Degree T C T 112 TD i i 4 148.8 i 4 *3 Coefficient of Performance i i i6 3.4 At the outlet 7 of the condenser 2, hot water of ca. 99 *C was discharged with a temperature difference of ca. whereas at the outlet 19 of the evaporator 14, cold water of 7 0 C was discharged from the inlet temperature of 12 0

C.

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 Heat Transfer Tube Wire Fin Corrugated Tube (Double-tube) Outer Tube (Diameter) Inner Tube (Diameter) Length Heat Transfer Area Corrugation Pitch OD t ID 31.8 x 1.6 x 30.2 mm 19.05OD x 0.95 x 17.15ID mm 3 520 m x 4 2 0.84 m 7.2 mm o" miiiii I 1W I I I- Corrugation Depth Height of Fins Fin Pitch 19 0.31 mm 0.8 mm 0.48 mm Table 6 Condenser Table 6 .4~ o4 44 44 an ai 0 4 o 0 4 4 Exchange Heat Quantity (Kcal/h) Condenser Inlet Temp. of Water Condenser Outlet Temp. of Water

(OC)

Saturation Temp. Superheating Degree (oC) Supercooling Degree (oC) 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 Super- Saturation Superheating Zone cooling Zone Zone 13,630 20.4 96.2 84.6 50.6 46.6 180 303.9 6470 3370 18.7 3790 21.1 36.0 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.

-r 1 i t 0 20 Table 7 This Invention\ State A B C D E F Temperature OC 135.2 84.6 84.6 38.0 0.49 30.1 Pressure kgf/cm 2 25.6 25.6 25.6 25.6 3.2 3.2 Enthalpy kcal/kg ii i2 i 3 i4 5 i6 153.8 142.7 121.4 108.9 108.9 141.0 Conventional State a b c d e f Temperature oC 135.2 84.6 84.6 79.6 0.49 30.1 Pressure kgf/cm 2 25.6 25.6 25.6 25.6 3.2 3.2 Enthalpy kcal/kg i'1 i'2 i'3 i'4 i'5 i'6 153.8 142.7 121.4 119.9 119.9 141.0 Notes: 4* 4 4 '4 0 94 04 04 0l 9 04 41i 41r 9 4 4 The symbols A to whereas the symbols a Mollier chart to Fig.

system.

F designate the states of Fig. 4b to f designate corresponding states of 5b in the conventional heat pump From Table 7 above, the following values of performances are calculated.

Supercooling Temperature COP Degree Effectiveness This Invention 46.6 0 C 72.6% Conventional 5°C 7.8% 2.6 In this way, hot water of ca. 96 0 C was available with a temperature difference of ca. 76°C and cold water of 7°C was available with a temperature difference of 5 0

C.

eMbodiMe'r Thus far described, this invention provides a heat pump system having such a heat pump cycle that in the condenser, 01 6- r I I 4 rI 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 Sdifference 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 120 0 C available S. instantaneously. Moreover, it is possible to lessen the 4 capacity of the compressor because the flow rate of refrigerant can be decreased, thus curtailing the installation cost and running cost.

Claims

1. A heat pump system comprising a heat pump cycle of refrigerant including a compressor, a condenser, an expansion valve and an evaporator 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 operated, when the system is run, so that while conducting heat exchange 10 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 O 4inlet 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, 0o wherein said evaporator further includes a once-through p '.th I 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, 0 wherein said evaporator is constructed of a heat exchanger l of thorough counterflow type between the refrigerant and the a I 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 100'C. kz~. -I 1' rl r a-9 LI i 23 The heat pump system as set forth in claim i, wherein said condenser permits to discharge hot fluid having an outlet temperature higher than an inlet temperature of the fluid by 50 0 to 90 0 C.

6. The heat pump system as set forth in claim i, 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 an 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 for high boiling refrigerant including a compressor, a condenser and an expansion valve; a low-temperature side heat pump cycle for 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 high-temperature side heat pump cycle and the compressor and expansion valve of the low-temperature side heat pump 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 l.quefied and condensed and then o 0 000 o 0o 0 000 0 00 0 00 0 o~ 0 0 00B 0 00a 0 00 010 0 OD 0 o 0. 0 00 Or 0 09 00 0a 0 0b 0 004 I 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. 1 0 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 heat pump 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 3 0 the fluid by 30 to 100'C.

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

14. The heat pump system as set forth in claim 9, U1IiUl L.ILltA'Z11 -C v~c l i _I 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 900C. A heat pump system substantially as hereinbefore described and illustrated with reference to the accompanying drawings. DATED THIS 19TH DAY OF FEBRUARY, 1991. 1) OSAKA PREFECTURE 2) NISHIYODO AIR CONDITIONER CO., LTD. By Their Patent Attorneys: GRIFFITH HACK CO. Fellows Institute of Patent Attorneys of Australia. o0 4 i; i. I44 I 4 4 C-