WO2002044632A1

WO2002044632A1 - Variable capacity refrigerant-sourced heat pump

Info

Publication number: WO2002044632A1
Application number: PCT/AU2001/001540
Authority: WO
Inventors: Ronald David Conry
Original assignee: Turbocor Inc
Priority date: 2000-12-01
Filing date: 2001-11-28
Publication date: 2002-06-06
Also published as: MY138533A; AU2002223297A1; AUPR185700A0; TWI225918B

Abstract

A refrigeration system has a plurality of heat exchange units (27) with high and low pressure gas and liquid refrigerant lines (14,31,32,33) connected to a compressor (12), a condenser (19), and a secondary evaporator (24). A primary refrigerant discharge line (14) extends between the compressor (12) and each heat exchange unit, and a secondary discharge line (18) communicates with the primary discharge line (14) and the secondary condenser (19). A pressure balancing valve (21) is located in the secondary discharge line (18). Expansion valve means (34) are provided in the high pressure gas line (33) at each heat exchange unit (27), and valve means (28) at each heat exchange unit (27) control the flow of gas or liquid refrigerant from the high pressure liquid, high pressure gas and low pressure gas lines (14, 32,33) through the heat exchange units (27) according to sensed load demand.

Description

VARIAB E CAPACITY REFRIGERANT-SOURCED HEAT PUMP Field of the Invention

This invention relates to a variable capacity refrigerant-sourced heat pump and relates particularly to a heat pump adapted for use in providing air conditioning to a building or other structure. However, it will be appreciated that the features of the invention may also be applicable to heating and or cooling in other applications such as in a manufacturing process line, a materials handling system or in any other environment. However, for simplicity of description, the invention will be described with particular reference to its use in a multi-zone air conditioning system.

Background of the Invention

Many difficulties exist with the use of refrigeration systems for heat pump applications, and particularly those using relatively larger types of refrigeration compressors. The pumping capacity control of refrigeration systems using traditional reciprocating and screw compressors is inflexible. Many such compressors can only be unloaded in one or two stages, if at all. In other systems, unloading of the system may be effected by switching on or off one or more compressors in the system. The amount that a system may reduce capacity in an efficient manner is conditional upon the design of the compressor and the design of the system and its ability to operate under light load conditions.

One factor effecting the ability to control the refrigeration system capacity is the management of lubricating oil in the refrigeration system. Standard refrigeration compressors require lubricating oil and this oil is carried throughout the system with the refrigerant. Systems must therefore be designed to enable the oil to be trapped and conveyed back to the compressor. To allow for such oil return, systems need to be designed with minimum suction and liquid line diameters so that fluid velocity is sufficiently high to carry oil with the refrigerant.

With rninimized pipe sizes, line pressure losses are large thus compromising system efficiencies. If the oil return is not managed correctly, the compressor can starve for lubricant and may well seize in severe lubrication starvation situations. The cost of managing the oil circuity in an air conditioning or refrigeration system is often very expensive and results in system designs which prevent low load operation.

Description of the Prior Art

In most larger air conditioning installations, where centrifugal compressors are involved, the compressors are limited to being used in systems involving water cooled chillers. In some cases, however, air cooled chillers may also be used with centrifugal compressors. However, it has not been an option with known systems to utilize centrifugal compressors in heat pump applications or direct expansion systems. The oil reclaim problems of centrifugal compressors operated at low load for long periods of time have generally mitigated against their use.

Further, in most large air conditioning installations, multiple zones are required to cope with the varying loads in different parts of a building during different times of the day and during different times of the year. Thus, it is quite common for some areas or zones of a building to require heat whilst other areas or zones need cooling, and for further areas or zones to require neither heating nor cooling. These load requirements will vary throughout the day and from season to season.

The heating and cooling requirements for multiple zones can be handled in a number of ways. The most common is to use chilled water pumped through air handling units for cooling designated zones, and for using hot water to be pumped through a separate coil in the same air handling unit. Chilled water systems are most common because of the lower installation costs, but because a secondary fluid is used between the refrigeration system and the air, efficiencies are compromised. Another system uses water source heat pumps which give multiple zone control. This is done by using a number of separate refrigeration heat pump units with each zone having its own unit. The units are linked together by a common water circuit loop which is maintained at approximately 22°C. The units which are on a cooling cycle reject heat into the loop, and the units that are on a heating cycle take heat from the loop. For any excess heat in the loop, the heat is rejected through a cooling tower, and if addition heat is required for the loop this is added through a hot water boiler. This system, however, requires a large multiple of separate heat pump units which increases the installation costs and maintenance.

United States Patent No. 4016657 discloses a co-ordinated heat pump system, developed particularly for use in a freeze-drying application, which extracts heat from a heat source at a lower temperature and dehvers heat to a heat sink at a higher temperature, with both heat exchange operations performed at a desired rate. However, this heat pump system uses a cooling means to control the ratio of heat delivered to the heat sink and heat withdrawn. The system is not designed for application where different areas or zones of a building, for example, require different heating or cooling rates.

It is therefore desirable to provide an improved variable capacity heat pump which obviates at least some of the disadvantages of existing systems.

It is also desirable to provide a refrigeration system for multiple zones which is relatively simple and economic to manufacture and install. It is also desirable to provide an improved variable capacity refrigerant sourced heat pump system which enables individual temperature control of the separate zones.

It is also desirable to provide an improved refrigeration system or air conditioning system able to use a centrifugal compressor with infinite capacity control.

It is also desirable to provide an improved air conditioning system which is efficient in operation and is easy to control.

According to one aspect of the invention there is provided a refrigeration system including a compressor, a plurality of heat exchange units, a secondary condenser, a secondary evaporator, a primary refrigerant discharge line between the compressor and each heat exchange unit, a secondary discharge line communicating with the primary discharge line and the secondary condenser, said secondary discharge line having a pressure balancing valve, a high pressure liquid Hne between the secondary condenser and each heat exchange unit, a high pressure gas hne and a low pressure gas Hne between each heat exchange unit and the secondary evaporator, expansion valve means in the high pressure gas hne at each heat exchange unit, a reversing valve means at each heat exchange unit in each of the said discharge, high pressure Hquid, high pressure gas and low pressure gas Hnes, whereby each reversing valve means controls the flow of Hquid or gaseous refrigerant through the respective heat exchange units according to sensed load demand.

Preferably, the compressor is of the type described in our AustraHan Patent No 686174, being a high speed centrifugal compressor running on oilless bearings and incorporating a control system that enables infinitely variable capacity control ranging from about 10% to 100%. The use of such a compressor aHows even matching of load requirements of the system so that close control can be maintained and system flexibihty and efficiencies increased. The compressor may be a single stage compressor although, in a preferred form of the invention, the compressor is a two stage compressor. With the two stage compressor, a two stage expansion tank is connected to the secondary condenser through a first stage expansion device, such as a float valve, a gas outlet from the expansion tank being connected into the compressor between the two compressor stages. The expansion tank is also connected to the secondary evaporator through a secondary thermostatic expansion valve.

The heat exchange units are preferably in the form of reverse cycle fan coil units which can act either as primary condensers or primary evaporators, depending on the operation of the reversing valves. The air which passes through the fan coil units may therefore be cooled or heated as required, or if no heating or cooling is required, the fan coil units can be used solely for air recirculation.

A hot water boiler may be used in conjunction with the secondary evaporator to ensure that the refrigerant gas has a minimum of super heating to ensure that damage is not caused to the compressor from Hquid being carried thereto. Similarly, a cooHng tower may be required to supplement the operation of the secondary condenser. However, it will be appreciated that an exchange of heat between the secondary condenser and the secondary evaporator may be effected to reduce the need for external heating or cooHng. In order that the invention is more readily understood, an embodiment thereof wiU now be described with reference to the accompanying drawings. Description of the Drawings

Fig 1 is a schematic illustration of a refrigeration circuit in accordance with an embodiment of the invention and in which the system is operating primarily on a cooHng cycle;

Fig 2 is the circuit of Fig 1 but showing its operation on a heating cycle; and

Fig 3 is a circuit similar to that of Fig 1 but showing its operation in heating, cooHng and recirculating mode. Description of the Preferred Embodiment

Referring to Fig 1, the system of this embodiment comprises a compressor 12 having a gas inlet 13 and a high pressure primary gas discharge outlet Hne 14. The compressor 12 has a first stage 16 and a second stage 17 and runs on oilless bearings so that the refrigerant system design can be optimized without the need to compromise for the presence of oil.

The primary gas discharge Hne 14 is connected through a secondary discharge Hne 18 to a secondary condenser 19 through a pressure balancing valve 21. The secondary condenser 19 is connected to a Hquid receiver or two stage expansion tank 22 through a first stage expansion device 23, such as a ball valve. The expansion tank 22 is connected to a secondary evaporator 24 through a secondary thermostatic expansion valve 26.

The system of this embodiment incorporates a pluraHty of zones in each of which is a heat exchanger in the form of a fan coil unit 27 which can act either as a primary condenser or as a primary evaporator for the system. Alternatively, when a particular fan coil unit is not being used for heating or cooHng, it operates as a recirculating fan.

Associated with each fan coil unit 27 is a reversing valve 28. The reversing valves may either switch on or off the various Hnes to which they are connected or they may throttle individual Hnes, depending on the type used and the design criteria of the system.

As shown in Fig 1, the primary gas discharge Hne 14 connects to each of the fan coil units 27 through the gas Hne 29, and a Hquid return Hne 31 returns Hquid refrigerant to the secondary condenser 19. The secondary evaporator 24 is also connected to each fan coil unit 27 through the suction Hne 32 and the reversing valves 28 as is the high pressure Hquid Hne 33. The high pressure Hquid Hne 33 is also connected to each fan coil unit 27, through a respective reversing valve 28, and to the outlet side of the expansion tank 22. The high pressure Hquid Hne 33 provides high pressure Hquid refrigerant from the secondary condenser 19 and expansion tank 22 to the fan coil units 27 through expansion devices 34 located between the respective reversing valve 28 and the respective fan coil unit 27.

During the cooHng cycle as illustrated in Fig 1, when all of the zones have a full cooHng load, such as on a hot summers day, the hot gas is pumped from the compressor 12 as a super heated gas into the primary high pressure discharge line 14. As all of the reversing valves 7 are switched to "cooHng", the valves in the gas Hne 29 are closed as are the valves in the Hquid return Hne 31. Therefore, the pressure in the primary gas discharge Hne 14 rapidly rises and the pressure balancing valve 21 opens to enable the high pressure gas to pass into the secondary condenser 19. The heat is removed from the gas in the condenser, possibly by using a cooHng tower 36. The gas condenses into a high pressure Hquid_. which is fed into the expansion tank 22 through a first stage expansion device 23. The pressure of the Hquid is reduced to approximately half way between the condensing pressure and the evaporating pressure. Because the Hquid has dropped in pressure, the temperature of the liquid has also dropped, and the heat that was in the Hquid flashes off some of the Hquid into gas. This gas is fed back into the intermediate stage between the first stage 16 and second stage 17 of the compressor along the Hne 37. The intermediate pressure Hquid in the expansion tank is fed to the various expansion devices 34 through the high pressure Hquid Hne 33 and the opened reversing valves 28 in that Hquid Hne 33. As the liquid refrigerant passes through the fan coil units 27 heat is taken in from the air circulating through the units and gasifies the refrigerant which passes along the suction Hne 32 into the secondary evaporator 24. As the loads of the individual zones vary, the amount of gas actually pumped also varies. The compressor control 38 receives signals relating to temperatures and pressures at various parts of the system and can determine the total load on the system at any given time. The compressor speed is then able to be controUed along with internal guide vanes 39 in the gas inlet 13 to thereby vary the refrigeration capacity to match the load. Further control of the system may be effected by varying the capacity of the fan in each fan coil unit 27 through a variable speed control or the like.

Because the total system is dynamic at all times, the actual pumping capacity requirements of the compressor will be constantly varying. However, use of a

^' high speed, centrifugal compressor running on oilless bearings such as described in our AustraHan Patent No. 686174. Such a compressor is able to operate with high efficiency over a range of pumping capacity.

Referring to Fig 2, when aU the zones have a full heating load, such as on a cold day, the gas pumped from the compressor 12 passes into the high pressure primary gas discharge Hne 14 as a super heated gas. As all of the reversing valves 28 are switched to the heating mode, the discharge gas is pumped directly into the fan coil units 27 which now act as condensers. The high pressure gas condenses into a high pressure liquid thereby giving off heat which is transferred to the air circulation through the fan coil units 27. The high pressure Hquid passes along the liquid return Hne 31 to the secondary condenser 19 where further sub-cooling may take place to increase the performance of the system. From the secondary condenser, the Hquid is fed to the secondary evaporator 24 through the first stage expansion device 23, expansion tank 22 and secondary thermostatic expansion valve 26.

As previously described, any gas flashed off in the expansion tank 22 is fed back to the intermediate stage of the compressor 12 between the first stage 16 and the second stage 17 along the Hne 37. As in the cooHng cycle, if a variable air volume system is required to more evenly match the load in any particular zone, a variation of the speed of the fan in the fan coil unit 27 enables the capacity of the individual fan coil units to be varied. The Hquid in the secondary evaporator 24 evaporates to a gas, and a boiler

41 may be used to provide heat to the secondary evaporator 24 to ensure that the gas entering the compressor 12 is at a minimum super heat.

As the loads in the different zones vary, the amount of gas actuaUy pumped also varies accordingly. When the condition in a particular zone is satisfied, the reversing valve 28 in that zone wiU go into a neutral position and the refrigerant will cease to flow through that particular unit 27. As previously indicated, the reversing valves may be constructed to vary the flow of refrigerant or to switch either one or off.

Referring to Fig 3, when the various zones have different loadings, such as in Spring or Autumn some zones may require heating while others require cooHng and others will be satisfied. In these cases, which accounts for much of the year in a lot of climates, the gas is pumped from the compressor 12 as a super heated gas into the primary gas discharge Hne 14. As some of the reversing valves 28 will be calHng for cooHng and others for heating, and others turned off, the gas in the gas discharge Hne 14 is pumped into those units which are calHng for heating. The heat is removed from the gas and the gas condenses into a high pressure Hquid thereby transferring heat to air passing through the appropriate fan coil unit 27. The Hquid passes along the Hquid return Hne 31 to the secondary condenser 19 and further condensing and sub-cooHng takes place. From here, the sub-cooled Hquid is fed through the expansion tank 22 and associated first stage expansion device 23 and then through the Hquid Hne 33 to those fan coil units 27 which are calling for cooling. The Hquid passes through the primary expansion devices 34 through the respective fan coil units 27 where the Hquid takes in heat and becomes gaseous and returns along the suction Hne 32 to the secondary evaporator 24. In the event that there is insufficient condensing taking place in a particular fan coil unit 27, surplus gas is fed to the secondary condenser and the heat is rejected through the cooHng tower 36. Similarly, heat may be provided by a boiler 41 to the secondary evaporator 24 to ensure that gas passing from the secondary evaporator 24 has πύnimum super heat. However, it wiU be appreciated that heat may be passed between the secondary condenser and secondary evaporator, as necessary to reduce the need for external heating and cooHng sources.

Many modifications may be made in the design and/or construction of a variable capacity refrigerant sourced heat pump in accordance with the present invention and all such modification which come within the scope of the invention shall be deemed to be within the ambit of the above description.

Claims

1. A refrigeration system including a compressor, a pluraHty of heat exchange units, a secondary condenser, a secondary evaporator, a primary refrigerant discharge Hne between the compressor and each heat exchange unit, a secondary discharge line communicating with the primary discharge Hne and the secondary condenser, said secondary discharge line having a pressure balancing valve, a high pressure Hquid line between the secondary condenser and each heat exchange unit, a high pressure gas Hne and a low pressure gas line between each heat exchange unit and the secondary evaporator, expansion valve means in the high pressure gas Hne at each heat exchange unit, valve means at each heat exchange unit in each of the said discharge, high pressure Hquid, high pressure gas and low pressure gas Hnes, whereby each valve means controls the flow of Hquid or gaseous refrigerant through the respective heat exchange units according to sensed load demand.

2. A refrigeration system according to claim 1 wherein the compressor is a high speed centrifugal compressor running on oilless bearings.

3. A refrigeration system according to claim 1 or claim 2 including a control system which provides variable capacity control of the compressor in accordance with load on the system.

4. A refrigeration system according to any one of the preceding claims wherein the compressor is a single stage compressor.

5. A refrigeration system according to any one of claims 1 to 3 wherein the compressor is a two stage compressor.

6. A refrigeration system according to claim 5 wherein an expansion tank is connected to the secondary condenser through a first stage expansion device, a gas outlet from the expansion tank being connected into the compressor between the two compressor stages.

7. A refrigeration system according to claim 6 wherein the expansion tank is also connected to the secondary evaporator through a secondary thermostatic expansion valve.

8. A refrigeration system according to any one of the preceding claims wherein the heat exchange units are in the form of reverse cycle fan coil units which can act either as primary condensers or primary evaporators, depending on the operation of the valve means.

9. A refrigeration system according to any one of the preceding claims wherein each heat exchange unit may operate as a primary condenser or primary evaporator independently of the operation of other heat exchange units in the system.

10. A refrigeration system according to any one of the preceding claims wherein if no heating or cooHng is required, the fan coil units can be used solely for air recirculation.

11. A refrigeration system according to any one of the preceding claims wherein a hot water boiler may be used in conjunction with the secondary evaporator to ensure that the refrigerant gas has a minimum of super heating.

12. A refrigeration system according to any one of the preceding claims wherein a cooling tower is used to supplement the operation of the secondary condenser.

13. A refrigeration system according to any one of the preceding claims wherein an exchange of heat between the secondary condenser and the secondary evaporator is effected to reduce the need for external heating or cooHng.

14. A refrigeration system according to any one of the preceding claims wherein the valve means comprises a reversing valve operable to select one of a pluraHty of operating modes whereby to control the flow of Hquid or gaseous refrigerant to the heat exchange units.

15. A refrigeration system substantially as hereinbefore described with reference to the accompanying drawings.