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APPARATUS, METHOD AND SOFTWARE FOR

USE WITH AN AIR CONDITIONING CYCLE

Title:

European Patent EP1540140

http://www.freepatentsonline.com/EP1540140.html

Document Type and Number:

Link to this page:

Abstract not available for EP1540140Abstract of corresponding document: WO2004022920A turbine (21) for generating power includes a rotor (23) in a rotorchamber and at least one nozzle (22) for supplying a fluid to drivethe rotor (23). The flow of the fluid from the nozzle exit (12) isperiodically interrupted by at least one flow interrupter means (7,11), thereby raising a pressure of the fluid inside the nozzle (22).Two such turbines (21) could be used in a thermodynamic cycle;the first turbine located downstream of a compressor and upstreamof a heat exchanger and the second turbine located downstream of

an evaporator and upstream of the compressor.

Abstract:

Page 1 of 31APPARATUS, METHOD AND SOFTWARE FOR USE WITH AN AIR CONDITIONI...

8/18/2007http://www.freepatentsonline.com/EP1540140.html

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DRYSDALE, KENNETH WILLIAM PATTEEVES, PAUL THOMASCASEY, ROBERT THOMAS

Inventors:

EP20030793477Application Number:

09/05/2003Filing Date:

06/15/2005Publication Date:

View Patent Images:Images are available in PDF form when logged in. To view PDFs,

View patents that cite this patentReferenced by:

Click for automatic bibliography generationExport Citation:

DRYSDALE KENNETH WILLIAM PATTE (AU)Assignee:

F01D1/02BPrimary Class:

F01D1/08, F25B11/02Other Classes:

F01D1/08, F01D25/00, H02P9/14, H02K17/42, F25B11/02,F25B1/04, F25B1/053

International Classes:

CLAIMS 1. A turbine for generating power including: a rotorchamber; a rotor rotatable about a central axis within said rotorchamber; at least one nozzle including a nozzle exit for supplying afluid from a fluid supply to said rotor to thereby drive said rotorand generate power; at least one exhaust aperture to, in use,exhaust said fluid from said turbine; wherein the flow of said fluidfrom said at least one nozzle exit is periodically interrupted by atleast one flow interrupter means, thereby raising the pressure ofsaid fluid inside said at least one outer nozzle.

2. The turbine of claim 1 including at least one fluid storage meansbetween said fluid supply and said at least one outer nozzle.

3. The turbine of claim 1 or claim 2, wherein said at least one flowinterrupter means substantially stops the flow of said fluid fromsaid at least one nozzle exit until the pressure inside said at leastone nozzle rises to a preselected minimum pressure, which is lessthan or equal to the pressure of the fluid supply.

4. The turbine of any one of claims 1 to 3, wherein in use, saidflow of said fluid from said at least one nozzle is interrupted bysaid at least one interrupter means for a period sufficient to bring

Claims:



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said fluid immediately upstream of said at least one outer nozzlesubstantially to rest.

5. The turbine of any one of claims 1 to4, wherein said rotor has aplurality of channels shaped, positioned and dimensioned toprovide a turning moment about said central axis when refrigerantfrom said at least one nozzle enters said channels.

6. The turbine of any one of claims 1 to 4, wherein said rotor is hasa plurality of blades shaped, positioned and dimensioned toprovide a turning moment about said central axis when refrigerantfrom said at least one nozzle contacts said blades.

7. The turbine of any one of claims 1 to 6, wherein said at least oneflow interrupter means includes at least one vane connectable toand moveable with an outer periphery of said rotor and adapted to

interrupt the flow of said fluid out of said at least one outer nozzleexit when said at least one vane is substantially adjacent said atleast one nozzle exit.

8. The turbine of claim7, wherein said flow interrupter meansincludes a plurality of said vanes substantially evenly spaced apartaround said outer periphery of said rotor.

9. The turbine of any one of claims 1 to 8, when included in a heatpump circuit, wherein said fluid supply is a positive displacementcompressor.

10. The turbine of claim 9 when dependent on claim 2, whereinsaid fluid storage means has a capacity at least equal to adisplacement of said positive displacement compressor.

11. The turbine of claim 9 or claim 10, wherein said at least oneexhaust aperture includes diffuser and expander sections todecrease the velocity of said fluid and maintain the pressure of thefluid low once it has decelerated to a subsonic velocity.

12. The turbine of any one of claims 9 to 11, wherein said at leastone nozzle in use supplies said fluid to said rotor at a sonic or

supersonic velocity.

13. A method of communicating a fluid supplied by a fluid supplymeans at a fluid supply means pressure to a turbine rotor, themethod including; providing at least one nozzle for communicatingsaid fluid from said fluid supply means to said turbine rotor, tothereby drive said rotor, the method further including providing atleast one flow interrupter means to



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periodically interrupt the flow of said fluid out of said at least onenozzle, thereby raising the pressure of said fluid inside said at leastone nozzle to a preselected minimum pressure which is less orequal to said fluid supply means pressure before resuming the flowof said fluid out of said at least one nozzle.

14. The method of claim 13, wherein said preselected minimumpressure is sufficient to cause the fluid to reach the local sonicvelocity at a throat of the nozzle.

15. The method of claim 14, including accelerating fluid exitingsaid at least one nozzle to supersonic velocities.

16. A turbine including a rotor including two or more spaced apartrotor windings and a stator including a plurality of stator windingsabout said rotor, wherein at least two of said stator windings areconnected to a controllable current source, each controllable

current source operable to energise the stator windings to which itis connected.

17. The turbine of claim 16, wherein each controllable currentsource is operable to energise the stator windings to which it isconnected after the rotor has reached a predetermined velocity.

18. The turbine of claim 17, wherein the predetermined velocity isthe terminal velocity for the current operating conditions of theturbine.

19. The turbine of any one of claims 16 to 18, wherein each currentsource increases or decreases the current through their respectivestator windings dependent on a measure of the power output fromthe stator windings.

20. A method of controlling a turbine including a rotor includingtwo or more spaced apart rotor windings and a stator including aplurality of stator windings about said rotor, wherein at least twoof said stator windings are connected to a controllable currentsource, each controllable current

source operable to energise the stator windings to which it isconnected, the method including repeatedly measuring the poweroutput from the stator windings and increasing the current throughthe windings if the current measure of power output is greater thana previous measure of power output and decreasing the currentthrough the windings if the current measure of power output is lessthan a previous measure of power output.

21. A thermodynamic cycle including a compressor, a first turbine



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downstream of the compressor, a heat exchanger locateddownstream of the first turbine and operable to reject heat from thecycle to another thermodynamic cycle, an evaporator downstreamof the heat exchanger and a second turbine downstream of theevaporator and upstream of the compressor.

22. A thermodynamic cycle including a compressor, a condenserdownstream of the compressor, a first turbine downstream of thecondenser, an evaporator downstream of the first turbine and asecond turbine downstream of the evaporator and upstream of thecompressor.

23. The thermodynamic cycle of claim 22 further including a heatexchanger located between said first turbine and said evaporator,the heat exchanger operable to reject heat to anotherthermodynamic cycle.

24. The thermodynamic cycle of any one of claims 21 to 23,

wherein the first and second turbines are turbines according to anyone of claims 1 to 11.

25. The thermodynamic cycle of any one of claims 21 to 24,wherein the first and second turbines are turbines according to anyone of claims 17 to20.

26. A control system for a thermodynamic cycle including acompressor, the control system including : sensing means forproviding a measure of an output of the thermodynamic cycle ;control means for the compressor, wherein the control means is in

communication with said sensing means to receive as inputs saidmeasure of an output of the thermodynamic cycle and a measure ofthe work input of the compressor; wherein the control means isoperable to compute a measure of efficiency from said inputs andvary the speed of the compressor to maximise said measure ofefficiency or to maintain said measure of efficiency at apredetermined level.

27. The control system of claim 26, further including second

control means for a TX Valve or equivalent and sensing means forproviding a measure of the temperature of a controlled area,wherein the second control means receives as a further input saidmeasure of the temperature of a controlled area, and is operable toopen or close the TX valve or equivalent in response to sensedvariations in temperature in the controlled area in relation to atarget measure.

28. The control system of claim 26 or claim 27, wherein the secondcontrol means further receives as an input a measure indicative of



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the amount of refrigerant in the cycle which is vaporised after anevaporation phase in the cycle and to open or close the TX valve orequivalent to maintain vaporised refrigerant after the evaporationphase.

29. The control system of any one of claims 26 to 28, wherein theoperation of the second control means to maintain vaporisedrefrigerant after the evaporation phase is performed after apredetermined delay from the control means opening or closing theTX valve in response to said sensed variations of temperature.

30. The control system of any one of claims 26 to 29 includingthird control means for a condenser in the thermodynamic cycle,the control system varying the operation of the condenser tomaintain a required level of cooling of refrigerant by thecondenser.

31. The control system of any one of claims 26 to 30, operable to

control a turbine according to claim 17 and including fourthcontrol means to control

the direct current through the stator windings of said turbine.

32. The control system of claim 31, operable control the directcurrent through the stator windings to dynamically maintain thebalance of said turbine when loaded.

33. The control system of claim 31, wherein the control means,second control means, third control means and fourth controlmeans is a single microcontroller or microprocessor or a pluralityofmicrocontrollers or microprocessors with at least selectedmicrocontrollers or microprocessors in communication with eachother to allow management of the timing of the functions of thecontrol system.

34. A method of controlling a turbine including a rotor includingtwo or more spaced apart rotor windings and a stator including aplurality of stator windings about said rotor, wherein at least twoof said stator windings are connected to a controllable current

source, each controllable current source operable to energise thestator windings to which it is connected, the method includingadjusting the current through the windings in order to dynamicallymaintain the balance of said rotor.

Description:



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APPARATUS, METHOD AND SOFTWAREFOR USE WITH AN AIR CONDITIONING CYCLETECHNICAL FiELDThe present invention relates to heat pumps, turbines for use withheat pumps and/or generators for use with heat pumps, and inparticular, but not exclusively, to improved refrigeration or airconditioning methods and apparatus and to turbines and/orgenerators for use therewith.

BACKGROUNDPresent refrigeration cycles reject heat to the atmosphere. In somecases a portion of the energy which would otherwise be rejectedmay be recovered from the cycle, thereby increasing the overallefficiency.

Figure 1 shows a diagrammatic representation of a heat pumpcircuit of the prior art. Hot, high pressure refrigerant liquid enters a

throttling device, often referred to as a Tx valve, which reduces itspressure and temperature at constant enthalpy. The heat absorbingvapour is passed through a heat exchanger or"evaporator"whichabsorbs heat from ambient temperature air blown across itssurfaces by a fan, cooling the air and thereby providing therefrigeration effect and causing it to expand. The acquisition ofheat causes the liquid to flash to vapour and expand.

The heat laden working fluid vapour is then passed into anaccumulator which has an internal structure designed to allow anyremaining liquid to boil off prior to entering the compressor.

The energy rich warm working fluid vapour enters a compressor,which as a result of a work input, compresses the vapour thusraising its temperature and pressure. A significant portion of thework input into the compressor re-appears as the heat ofcompression thussuperheating the working fluid vapour.

The superheated working fluid vapour thus has its temperature

elevated above that of the ambient temperature of the environmentand enters a condenser, which has a structure similar to that of theevaporator. A heat exchange then occurs between the superheatedworking fluid vapour and the environment which is at a lowertemperature. The heat exchange continues until sufficient heat isremoved from the working fluid to cause a change of state fromhot vapour to hot liquid.

The hot working fluid liquid enters a reservoir, usually referred toas a"receiver"which has a sufficiently large volume to support the



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requirements of the thermodynamic cycle and withstand the highpressure in the discharge line of the compressor. The hot highpressure refrigerant liquid then enters the TX valve to complete thethermodynamic cycle.

Air conditioning systems have become a huge draw on electricitypower in many of the major cities of the world and are viewed asan essential component of many large buildings in order tomaintain a level of environmental control within the building. Atthe same time as air conditioning systems continue to increase innumber, it is becoming increasingly recognised that electricity is alimited resource and in some places demand is exceeding supply oris forecast to in the near future.

It has become important to identify potential areas for saving inelectricity consumption. If any savings can be made in airconditioning systems, then there is potential to make an overallhuge saving in the consumption of electricity.

The saving of electricity can also lead to savings in powerdistribution infrastructure upgrades. Such upgrades are becomingnecessary to deal with increasing peak loads introduced by arapidly growing air conditioning market.

OBJECT OF THE INVENTIONIt is an object of a preferred embodiment of the invention toprovide apparatus for a heat pump and/or a heat pump which willincrease the utilization of available energy in such apparatus atpresent.

It is an alternative object of a preferred embodiment of theinvention to provide a method of controlling a heat pump whichwill increase the efficiency of such apparatus at present.

It is an alternative object of a preferred embodiment of theinvention to provide a method of controlling a turbine andgenerator which will increase the efficiency of such apparatus at

present.

It is an further alternative object of a preferred embodiment of theinvention to provide a turbineand/or a method of communicating afluid to a turbine which will increase the utilization of availableenergy from such fluid at present.

It is a still further alternative object to at least provide the publicwith a useful choice.



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Other objects of the present invention may become apparent fromthe following description, which is given by way of example only.

SUMMARY OF THE INVENTIONAccording to a first aspect of the invention there is provided aturbine for generating power including : a rotor chamber; a rotorrotatable about a central axis within said rotor chamber; at leastone nozzle including a nozzle exit for supplying a fluid from afluid supply to said rotor to thereby drive said rotor and generatepower; at least one exhaust aperture to, in use, exhaust said fluidfrom said turbine; wherein

the flow of said fluid from said at least one nozzle exit isperiodically interrupted by at least one flow interrupter means,thereby raising the pressure of said fluid inside said at least oneouter nozzle.

Preferably, the turbine includes at least one fluid storage meansbetween said fluid supply and said at least one outer nozzle.

Preferably, said at least one flow interrupter means substantiallystops the flow of said fluid from said at least one nozzle exit untilthe pressure inside said at least one nozzle rises to a preselectedminimum pressure, which is less than or equal to the pressure ofthe fluid supply.

Preferably, when the turbine is in use, said flow of said fluid fromsaid at least one nozzle is interrupted by said at least oneinterrupter means for a period sufficient to bring said fluidimmediately upstream of said at least one outer nozzlesubstantially to rest.

Preferably, said rotor has a plurality of channels shaped, positionedand dimensioned to provide a turning moment about said centralaxis when refrigerant from said at least one nozzle enters saidchannels.

Preferably, said rotor is has a plurality of blades shaped, positionedand dimensioned to provide a turning moment about said central

axis when refrigerant from said at least one nozzle contacts saidblades.

Preferably, said at least one flow interrupter means includes atleast one vane connectable to and moveable with an outerperiphery of said rotor and adapted to interrupt the flow of saidfluid out of said at least one outer nozzle exit when said at leastone vane is substantially adjacent said at least one nozzle exit.

Preferably, said flow interrupter means includes a plurality of said

Page 10 of 31APPARATUS, METHOD AND SOFTWARE FOR USE WITH AN AIR CONDITIO...


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vanes substantially evenly spaced apart around said outer peripheryof said rotor.

Preferably, said turbine is included in a heat pump circuit, whereinsaid fluid supply is a positive displacement compressor.

Preferably, said fluid storage means has a capacity at least equal toa displacement of said positive displacement compressor.

Preferably, said at least one exhaust aperture includes diffuser andexpander sections to decrease the velocity of said fluid andmaintain the pressure of the fluid low once it has decelerated to asubsonic velocity.

Preferably, said at least one nozzle in use supplies said fluid to saidrotor at a sonic or supersonic velocity.

According to a second aspect of the invention there is provided amethod of communicating a fluid supplied by a fluid supply meansat a fluid supply means pressure to a turbine rotor, the methodincluding; providing at least one nozzle for communicating saidfluid from said fluid supply means to said turbine rotor, to therebydrive said rotor, the method further including providing at least oneflow interrupter means to periodically interrupt the flow of saidfluid out of said at least one nozzle, thereby raising the pressure ofsaid fluid inside said at least one nozzle to a preselected minimumpressure which is less or equal to said fluid supply means pressurebefore resuming the flow of said fluid out of said at least onenozzle.

Preferably, said preselected minimum pressure is sufficient tocause the fluid to reach the local sonic velocity at a throat of thenozzle.

Preferably, the method includes accelerating fluid exiting said atleast one nozzle to supersonic velocities.

According to a third aspect of the present invention, there isprovided a turbine including a rotor including two or more spacedapart rotor windings and a stator including a plurality of statorwindings about said rotor, wherein at least two of said statorwindings are connected to a controllable current

source, each controllable current source operable to energise thestator windings to which it is connected.



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Preferably, each controllable current source is operable to energisethe stator windings to which it is connected after the rotor hasreached a predetermined velocity.

Preferably, the predetermined velocity is the terminal velocity forthe current operating conditions of the turbine.

Preferably, each current source increases or decreases the currentthrough their respective stator windings dependent on a measure ofthe power output from the stator windings.

According to a fourth aspect of the invention, there is provided amethod of controlling a turbine including a rotor including two ormore spaced apart rotor windings and a stator including a pluralityof stator windings about said rotor, wherein at least two of saidstator windings are connected to a controllable current source, eachcontrollable current source operable to energise the stator windings

to which it is connected, the method including repeatedlymeasuring the power output from the stator windings andincreasing the current through the windings if the current measureof power output is greater than a previous measure of power outputand decreasing the current through the windings if the currentmeasure of power output is less than a previous measure of poweroutput.

According to a fifth aspect of the invention, there is provided athermodynamic cycle including a compressor, a first turbinedownstream of the compressor, a heat exchanger locateddownstream of the first turbine and operable to reject heat from thecycle to another thermodynamic cycle, an evaporator downstreamof the heat exchanger and a second turbine downstream of theevaporator and upstream of the compressor.

According to a sixth aspect of the invention, there is provided athermodynamic cycle including a compressor, a condenserdownstream of the

compressor, a first turbine downstream of the condenser, an

evaporator downstream of the first turbine and a second turbinedownstream of the evaporator and upstream of the compressor.

Preferably, the thermodynamic cycle further includes a heatexchanger located between said first turbine and said evaporator,the heat exchanger operable to reject heat to anotherthermodynamic cycle.

Preferably, the first and second turbines are turbines according tothe preceding paragraphs.



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Preferably, the control system is operable to control the directcurrent through the stator windings to dynamically maintain thebalance of said turbine when loaded.

Preferably, the control means, second control means, third controlmeans and fourth control means is a single microcontroller ormicroprocessor or a plurality ofmicrocontrollers ormicroprocessors with at least selectedmicrocontrollers ormicroprocessors in communication with each other to allowmanagement of the timing of the functions of the control system.

According to an eighth aspect of the invention, there is provided amethod of controlling a turbine including a rotor including two ormore spaced apart rotor windings and a stator including a pluralityof stator windings about said rotor, wherein at least two of saidstator windings are connected to a controllable current source, eachcontrollable current source operable to

energise the stator windings to which it is connected, the methodincluding adjusting the current through the windings in order todynamically maintain the balance of said rotor.

Further aspects of the present invention, which should beconsidered in all its novel aspects, will become apparent from thefollowing description, given by way of example only and withreference to the accompanying drawings.

BRIEF DESCRIPTION OF DRAWINGSFigure 1: Shows a prior art thermodynamic cycle.

Figure 2: Shows a first thermodynamic cycle according to anaspect of the present invention.

Figure 3: Shows a second thermodynamic cycle according to anaspect of the present invention.

Figure 4: Shows a cross-sectional view of a first turbine accordingto an aspect of the present invention.

Figure 5: Shows a cross-sectional view of a second turbineaccording to an aspect of the present invention.

Figure 6: Shows an enlarged view of a channel of the turbine ofFigure 5.

Figure 7: Shows a third thermodynamic cycle illustrating a controlsystem according to an aspect of the present invention.

Figures 8-10,12 : Show flow charts of a method of controlling a



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thermodynamic cycle according to aspects of the present invention.

Figure 11: Shows a diagram of a generator according to an aspectof the present invention.

Figure 13: Shows a flow chart of an initialisation subroutine for thecontrol system.

Figure 14: Shows a flow chart of a schduling subroutine for thecontrol system.

BRIEF DESCRIPTION OF PREFERRED EMBODIMENTS OFTHE INVENTIONThe present invention is described herein with reference to its

application to a refrigeration cycle. Those skilled in the art willrecognise that the heat pumping circuit described may have avariety of uses, for example air conditioning, refrigeration orheating. Those skilled in art will also recognise that theterm"refrigerant"is used to describe any working fluid suitable foruse in such a circuit or cycle.

A simple refrigeration circuit of the prior art shown in Figure 1may include, in order, a compressor, a condenser, a receiver, athrottling valve (TX valve), an evaporator and an accumulator.Some embodiments of the prior art may combine two of theelements shown in Figure 1 into a single device, for example somecompressors may also include an accumulator, but the function ofeach element is usually present in the circuit.

The term"turbine"is used herein to describe a device whichconverts energy from a fluid stream into kinetic and/or electricalenergy. Those skilled in the art will appreciate that where theenergy is required in electrical form the turbine may include asuitable electric power generator or alternator.

Referring next to Figure 2 a heat pump apparatus of the presentinvention includes a first refrigerant circuit 10, which includes in

order, a first compressor 1 a condenser 8, a receiver 2, a TX valve,an evaporator 5 and a turbine 21. The turbine 21 converts energyfrom the refrigerant into kinetic and/or electrical energy, therebylowering the temperature and pressure of the first refrigerant. Ifrequired to result in a suitable density and pressure refrigerant forthe turbine, an expander (not shown) may be provided on one orboth of the upstream and downstream sides of the turbine 21.

In some embodiments the turbine 21 may be designed to avoidcooling the refrigerant to the point where drops of liquid



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refrigerant form within the turbine 21, as this may damage theworking surfaces within the turbine 21.

In alternative embodiments the turbine 21 may be adapted, forexample

through the use of appropriately robust materials to construct therotor blades, to allow condensation of the refrigerant withoutdamage to the turbine 21.

Those skilled in the art will appreciate which qualities of therefrigerant passing through the first evaporator 5 will affect theheat flow into the first evaporator 5. The refrigerant leaving thefirst evaporator 5 passes through a first accumulator 6 beforereturning to the first compressor 1. Those skilled in the art willappreciate that the receiver 2 and accumulator 6 provides the

refrigerant reservoirs for the circuit. The accumulator 6 is shown inoutline to represent the option that it forms a part of thecompressor 1.

Referring to Figure 3, an alternative heat pump according to thepresent invention is shown, which includes a first refrigerantcircuit 300 and second refrigerant circuit 400. In a preferredembodiment the second refrigerant cycle 400 may include anevaporator 405, accumulator, compressor, condenser, receiver andTX valve (not shown), arranged in the same order and performingthe substantially same function as a refrigeration circuit of the priorart. The second refrigerant may have a boiling point of less than10C, more preferably around0 C. A suitable second refrigerant maybe R22, R134A or R123, although those skilled in the art willappreciate that other refrigerants with suitably low boiling pointsmay be used.

The second refrigerant circuit 400 may be controlled by a controlsystem as described below with reference to Figure 7. If required,both refrigerant circuits may be controlled by a single controller.

In a preferred embodiment the temperature of the refrigerantentering the condenser of the refrigerant circuit 400 may be

above30 C, and preferably around60 C. The temperature of therefrigerant entering the evaporator of the refrigerant circuit 400may be at least10 C lower than the temperature of the refrigerantentering the condenser 304.

In some embodiments one or more thermoelectric generatorspositioned between a compressor and condenser may be providedin order to generate electricity. Thermoelectric generators may beparticularly useful if



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the refrigerant usedis R123, as the condensing temperature may beas high as180 C and the evaporation temperature between35 Cand10 C, thereby providing a large temperature differential.

The cycle 300 includes in clockwise order a compressor 301,condenser 307, first expander 302a, first turbine 302, secondexpander 302b, a heat exchanger 304, an evaporator 305 and asecond turbine 306.

The expanders may be included on both the input and output sidesof the turbine 302 to reduce the density of the working fluidentering the turbine 302, and to assist in maintaining a lowpressure at the output of the turbine 302 after the working fluidreturns to a subsonic velocity. In a preferred embodiment theexpander may ensure that there is no increase in the pressure of thefluid once it has decelerated to a subsonic velocity. Without an

expander the pressure at the turbine output would otherwise riseand impair the turbine performance.

Expanders (not shown) may also be included on one or both of theinput and output of the second turbine 306. The expanders willinclude a diffuser if the refrigerant is circulating at supersonicspeeds out of the turbine 306. The expanders on the inputs of theturbines 302,306 are necessary to lower the density of the workingfluid prior to entering the throat of the turbine nozzle. The lowerdensity will allow a larger throat size at the sonic point of theworking fluid and hence maintain a critical minimum mass flowrate so as to avoid any reduction in air conditioning efficiency.Ideally the mass flow rate should be the same as would beexperienced without the introduction of each turbine into thethermodynamic cycle. Thevolumetric expansion before the nozzletherefore lowers the density of the working fluid and allows alarger diameter nozzle throat to be used without impairing eitherthe subsonic/supersonic transition of the working fluid at the throator its mass flow rate.

In two further alternative cycles, one of either the refrigerant cycle400 and condenser 304 may be omitted.

Figure 4 shows a turbine 21, suitable for use with the heat pumpapparatus described in relation to Figures 1,2, 3. The turbine 21may also be used in a refrigerant circuit of the prior art, such as thecircuit shown in Figure 1 or in other refrigerant circuit, preferablyeither immediately upstream or downstream of the compressor,with expanders provided about the turbine 21 if necessary. The



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turbine 21 includes at least one outer nozzle 22 mounted in thehousing (not shown) of the turbine 21, which hasaconverging/diverging section adapted to accelerate the refrigerantflowing through it to sonic or supersonic speeds.

The turbine 21 is described below with reference to its use as partof a heat pump circuit, such as those described above, in which theworking fluid is refrigerant. The turbine 21 may perform thefunction of a TX valve in addition to generating power, allowing aTX valve to be omitted from the circuit. Those skilled in the artwill appreciate that other applications for the turbine 21 arepossible and that the working fluid may in these embodiments besome other suitable gaseous fluid.

The flow from each outer nozzle 22 is periodically interrupted byan interruption means. Two preferred interruption means areexplained below.

Those skilled in the relevant arts may be able to identify alternativemeans for interrupting the flow from an outer nozzle 22.

A first interruption means may include one or more vanes 7located proximate the outer periphery of the turbine rotor 23 andadapted to substantially prevent refrigerant from flowing from anouter nozzle 22 when the vane 7 is proximate the outer nozzleoutlet 12. Those skilled in the relevant arts will appreciate that thegap between the exit of the outer nozzle 22 and the vanes 7 isexaggerated in Figure 4 and that the actual gap will be smallenough to interrupt or significantly inhibit flow from the nozzle 22when the vanes 7 are adjacent the nozzle exit 12.

A second interruption means 11 may include an electronicallyoperated valve proximate the outer nozzle outlet 12. The secondinterruption

means 11 may have an extremely fast response and may, forexample, be similar in operation to an electronically operatedcommon rail diesel injector.

A refrigerant storage vessel 13 may be located proximate the outernozzle entrance 14. If the compressor supplying refrigerant to theouter nozzle 22 is a positive displacement compressor, then therefrigerant storage vessel 13 may have an internal volume at leastequal to a single displacement of the first compressor. Therefrigerant storage vessel 13 may have any capacity greater thanthe displacement of the compressor. The refrigerant storage vessel13 may preferably be an insulated spherical container located asclose as possible to the outer nozzle entrance 14.



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The vanes 7 and second interruption means 11 may stop the flowof refrigerant sufficiently rapidly to cause an adiabatic pressurerise in the outer nozzle 22 without a corresponding increase inenthalpy. The flow of refrigerant may be interrupted for a periodwhich is sufficiently long for the pressure inside the outer nozzle22, and more preferably inside the refrigerant storage vessel 13, toreach a preselected minimum pressure which is less than thepressure supplied by the first compressor. This pressure may beselected to ensure that when the vanes 7 and second interruptionmeans 11 are both open, the refrigerant exits the outer nozzle 22 atsonic or supersonic speeds.

The period of time that each vane 7 stops the flow from the outernozzle 22 depends on the circumference of the turbine rotor 23, therotational speed of the rotor 23 and the length of the vane 7 in thecircumferential direction. In some embodiments this period of timemay be sufficiently long that a second interrupter means 11 is notrequired.

In other embodiments the second interruption means 11 may becapable of closing sufficiently rapidly that the vanes 7 are notnecessary, but in many cases the vanes 7 may provide a relativelysimple interruption means, which is capable of closing the outernozzle outlet 12 at high speed.

The refrigerant storage vessel 13, vanes 7 and second interruptionmeans 11 may assist in increasing the amount of energy recoveredfrom the refrigerant while still allowing sufficient refrigerant toflow to provide an adequate overall heat absorption effect from arefrigerant circuit. This may facilitate or assist the omission of areceiver and TX valve from the refrigeration circuit.

The Applicant believes that when the interruption means closes,the mass flow of the working fluid, in this case refrigerant,between the outer nozzle 22 and the high pressure source feedingthe outer nozzle 22, which in most cases may be a first compressor,may decrease towards zero, and the pressure in the refrigerant

storage vessel 13 and outer nozzle entrance 14 may rise towardsthe maximum pressure of the discharge line of the firstcompressor. This upward pressure excursion is a function of thedecrease in mass flow rate of the fluid. When the mass flow rate iszero then the pressure difference across the outer nozzle 22 may besubstantially zero, therefore the pressure at the outer nozzleentrance 14 is at a maximum and the kinetic energy change in therefrigerant is zero and the enthalpy change is zero.

Thus, when the refrigerant is stopped the pressure rises at the outer



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nozzle entrance 14 to the maximum value provided by thecompressor and the enthalpy change is zero. The Applicant alsobelieves that if the period of time when the refrigerant isinterrupted is short in comparison to the time in which therefrigerant is allowed to flow, then the deterioration in overallmass flow in a refrigerant circuit of which the turbine 21 is acomponent will be minimal.

The Applicant further believes that an advantage of stopping themass flow through the outer nozzle 22 is that, if the period of theflow interruption is sufficiently short and the increase in pressureof the refrigerant occurs substantially adiabatically, there will beno change in the enthalpy of the stationary refrigerant in the outernozzle 22. Also, if the increase in internal energy during the timewhen the refrigerant is stationary and the refrigerant is compressedcompensates for the expansion of the refrigerant and its depletionof work during the time when the mass flow is flowing, which maybe achieved by properly selecting the ratio of time during which

the refrigerant flows to time in which the refrigerant is interrupted,then the enthalpy

extraction process may become substantially continuous. TheApplicant believes that this may result in an increased extraction ofenthalpy from the working fluid over systems of the prior art.

Those skilled in the art will also appreciate that the timing of thesecond interruption means 11 may be controlled by a processingmeans (not shown). The processing means may receiveinformation on the angular position of the turbine rotor 23 fromany suitable means, but preferably from a hall effect sensor orsimilar mounted on the turbine housing (not shown), which maysense a suitable index mark on the rotor 23. The processing meansmay also vary the speed of the turbine rotor 23 by varying theopening times of the second interrupter 11.

While the turbine rotor 23 is shown having an impulse type bladeconfiguration, the Applicant has found that interrupters asdescribed above are also particularly suited to use with other radialtype turbine designs, for example those used in automotive

turbochargers, as is shown in Figure 11.

Referring now to Figure 5, an alternative turbine rotor 23A isshown as having a plurality of substantially spiral shaped channels602 leading to a central exhaust aperture 603. The central exhaustaperture 603 may be central of the rotor 23A and may extendsubstantially in the direction of the central axis of the rotor 23A.The cross-sectional area of each channel 602 may continuouslydecrease between an inlet 604 and an outlet 605.



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Preferably the ratio of the area of the inlet 604 to the outlet 605may be substantially 6: 1 in order to promote hypersonic operationwith the minimum restriction to the flow of the working fluid.

Referring next to Figure 6, the centreline 606 of each channel 602may intersect a radius 607 of the rotor 23A on at least two points,608,609 between the inlet 604 and the outlet 605.

A fluid flow, represented by arrows F, may enter a channel 602through an inlet 604. As the direction of the fluid F is changedwithin the

channel 602 the change in momentum of the fluid F may result in aturning force on the rotor 23A. Preferably the turning force may betransmitted to either a suitable electrical energy generator or anyother suitable mechanism which may be powered by a rotating

shaft. It is preferred that the fluid F execute as close as possible toa180 change in direction within the channel 602 in order tomaximise the change in momentum and therefore the energyimparted to the rotor 23A.

The rotor 23A may be used with an electronic second interruptermeans as described above, although those skilled in the art willrecognize that the in some embodiments the spacing 610 betweenthe channel entrances 604 may act as an interrupter means.

Figure 7 shows an air conditioning/refrigeration cycle, generallyreferenced by arrow 100, according to another aspect of the presentinvention.

Like the cycle 300 shown in Figure 3, the cycle 100 may differfrom air conditioning or refrigeration cycles of the prior art in thatthe TX valve and receiver common to the cycles of the prior art,may be omitted. The TX valve is replaced by a turbine 114, whichin this embodiment is located between the condenser 105 andevaporator 122. An optional thermoelectric generator 103 mayprecede the condenser 105.

A second turbine 130 is placed between the output of evaporator

122 and the accumulator 128. Expander 130a and 130b if presentare placed about turbine 130. This is to ensure that the density ofthe working fluid entering turbine 130 is sufficiently low, so as toallow a sufficiently large diameter nozzle to be used within turbine130, without impairing the supersonic operation of 130, the massflow rate of the system or its cooling efficiency.

A secondary heat pump cycle referenced by arrow 200 contains aheat exchanger 201 which follows expander 114c and allows heatto be removed from the primary cycle 100, to ensure that the



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isolation circuit, which may only allow power to flow in onedirection, thus avoiding any feedback of mains power 150 to thegenerator 121.

Those skilled in the art will recognise that the air conditioningcycles described above may be more energy efficient than those ofthe prior art, due to energy recovered by the turbine and, whereused, the thermoelectric generator, as well as the control of thecompressor speed to optimize the overall Coefficient ofPerformance.

Figures 8 to 10 show a series of flow diagrams illustrating anexample of the computational process of the present invention thatmay be performed to control an air conditioning cycle, such as thecycles described herein in relation to Figures 1,2, 3,7, 8 or othercycles including those of the prior art if required.

The process may be controlled by any suitable microcontroller,

microprocessor or similar having a control output to control thedrive signal of a motor controller for a compressor. For clarity, inthe following description it is assumed that a microcontroller hasbeen used.

Referring to Figure 8, on power up or before execution of thecontrol algorithms, an initialisation routine may be performed inwhich selected flags, registers and counters may be initialised,typically by setting to zero if this is

required for the particular implementation of the controlalgorithms.

Referring to Figure 13, a flow chart illustrating a possibleinitialisation subrouting is shown. The time intervals at whichexternal devices (for example the compressor, TX valve,condenser, generator excitation) are serviced/optimised are enteredas DEL1 toDELn. For the particular heat pump that is beingcontrolled, a look-up table is determined and the entries for targetcoefficients of performance (COP3 toCOPn) for the heat pumpwhen operated at a specific temperature differentials across the

evaporator((T1-T3) (1) to(T1-T3) (n) ) are entered.

The microprocessor may read the state of a switch SW1. Theswitch SW1 dictates whether the microcontroller automaticallyschedules servicing/optimisation of the control parameters for theheat pump. The current state of any required flags, counters andregisters may also be read and then initialised.

A look-up table is then formed from the entered temperaturedifferentials(T1-T3) (1) to(T1-T3) (n) and their associated target



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coefficients of performance COP3 to COPn for use in theservicing/optimisation of the heat pump (see herein below).Finally, the microcontroller sets a flag that dictates manual orautomatic operation based on the status of the switchSW1.

The microcontroller receives as inputs the temperature of therefrigerant flowing into the evaporatorT1, the temperature of therefrigerant leaving the evaporator T2 and the compressor motorpower KW1. The set point for the heat load T3, the required motorspeed increment K2 and required motor speed decrement K3 forthe compressor and an air conditioning refrigerant constant K1 arealso entered. K1 may be determined experimentally for theparticular air conditioning cycle and represents the increment ofheat lifted per degree temperature change between T1 and T2.

Having received these inputs, the microcontroller then computesthe difference between T1 and T3. This difference is then used tolook up a corresponding coefficient of performance for the heat

pump in the stored look-up

table, where the coefficient of performance represents the heatlifted per unit work input.

In an alternative embodiment, instead of working to a target COP,the microcontroller may increase/decrease the compressor speed tomaximise the COP if the COP for the cycle does not justcontinually increase with compressor speed. Those skilled in therelevant arts will also appreciate that variables other than thetemperature difference across the evaporator may be used ifrequired.

If T1-T3 is less than or equal to zero, the heat pump is notoperating and nothing further is done by the microcontroller,which returns to the start of the algorithm. IfT1-T3 is greater thanzero, the actual coefficient of performance COP2, which is basedon the measured variablesT1, T2 and KW1 is computed accordingto equation 1:COP2 = K1IT1-T21/KW1 equation1Other measures relating the output of the cycle to the compressor

work input may be used if required. As herein described, thepresently contemplated preferred embodiment uses measures oftemperature difference to provide a measure of the useful heattransferred by the system, as temperature measurements may berelatively easily obtained. However, alternative measures ofsystem performance may be used that relate the system output tothe compressor input.

The computed co-efficient of performance COP2 is then comparedto the target coefficient of performance COP1. If the value of



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COP1 is less than COP2, the compressor speed is increased by K2.Conversely, if the target COP1 is greater than the computed COP2,the motor speed is decreased by K3. A delay subroutine (notillustrated) is then executed to allow for any lag in the response ofthe cycle to the change in compressor speed. The required timedelay can be determined experimentally by forcing adjustments ofthe compressor speed by increments of K2 and K3 and measuringthe maximum time for the air conditioning cycle to return to steadystate conditions. Any suitable delay subroutine may be used toachieve this delay. The delay

subroutine is completed after any control variable is changedbefore analysing and varying another control variable to ensurethat the system remains stable and/or to ensure that steady stateconditions are used to provide measures of the inputs to the controlalgorithms. The execution of control algorithms may be performed

periodically at predetermined time intervals, continuously with theappropriate time delay between each control cycle or on a schduledbasis.

Figure 9 shows diagrammatically a control algorithm to control theoperation of a TX valve, if one is provided in the heat pump. Thecontrol algorithm may also be applied to any controllable devicethat performs the same or similar function to a TX valve.

The microcontroller receives as temperature inputs the unsaturatedtemperature of the air exiting the evaporator T4 and a constant T5representing a superheat temperature value added to thetemperature of the working fluid at the evaporator output. It alsoreceives a pressure input P1 representing the pressure of theworking fluid at the evaporator output, a measurement of thecurrent status of a TX valve or equivalent TX1, and set steps K4and K5 for incrementing and decrementing the operation of the TXvalve respectively.

The microcontroller computes T6 as the sum of T4 and T5 andcomputes T7 as the product of P1 with a constant K6, whichfacilitates the conversion of pressure to temperature of the workingfluid. If the temperature T6 is less than T7, the TX valve is opened

by increment K4 and if the temperatureT6. is greater than T7, theTX valve is closed by increment K5.

Otherwise, the TX valve is maintained in its current position. Theincremental and decremental step size may optionally be the same(K4=K5). A delay subroutine is then executed in order to allow thecycle to reach a steady state or near steady state before any furtheraction is taken.

With variation of the TX valve setting, it may be advantageous to



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check that the TX valve is still operating so that the refrigerant inthe suction line of the compressor after the evaporator issufficiently super heated to be at

the vapour state. Therefore, each time when the delay subroutinefollowing variation of the TX valve is invoked, the microcontrollermay perform an additional check on the operation of the TX valve.This check may only be necessary if the control over the limits ofoperation of the TX valve is not already present as part of the TXvalve and if the existing control algorithms do not bound the TXvalve within an acceptable operating range.

With variations in the compressor speed and TX valve opening, theoperation of the condenser will also vary. Therefore, the controllermay also control the drive fan to a condenser. This process isshown in Figure 10.

The temperature inputs to the algorithm are T1 and T3 as definedherein above, the liquid line temperature T8, measured at apredetermined point in the heat pump, typically at a pointimmediately following the condenser and the target temperaturefor the liquid line temperature T10. The step size for an incrementin condenser fan speed K7 and step size for an increment incondenser fan speed K8 are also inputs to the algorithm togetherwith the current condenser fan speedCFS1, minimum condenserfan speed CFSmin and maximum condenser fan speed CFSmax.Although the steps that use CFSmin and CFSmax are notillustrated in Figure 11 the values of CFSmin and CFSmax boundthe allowable speed of the compressor fan.

The microcontroller first calculates T11 as the difference betweenT3 and T1 and terminates the control algorithm for the condenserfan speed if T3 is greater than or equal toT1. If T3 is less thanT1,the cycle is operational and heat extracted by the condenser.Themicrocontroller then calculates T12 as the difference ofT10and T8 and if the target temperatureT10 is less than the actualtemperature T8 the current compressor speed CFS1 is increased byK7 and if T10 is greater than T8 the current compressor speed isdecreased by K8. A further time delay is invoked after variation of

the condenser fan operation.

The microprocessor may also vary the timing of the secondinterrupter 11 to optimize a selected parameter of each refrigerantcircuit. In some embodiments the heat absorbed by an evaporatormay be the selected

parameter, while in other embodiments the total power input one



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or more of the compressors may be the selected parameter.

Figure 14 shows diagrammatically a control algorithm for thescheduling of control/optimisation algorithms described hereinabove. A table of time parameters is stored in memory, whichspecifies when each algorithm is to be executed. This table of timeparameters will be entered by the heat pump administrator. Onpower up, a pointer is set to an initial value in the table of timeparameters and the clock started. The table of time parameters listssequentially all of the control algorithms, a time delay variable thatindicates the time delay that should occur between each executionfor that control algorithm and an address indicating where thecontrol algorithm can be found in memory.

The microcontroller reads the current time of the real time clockand adds the time delay indicated in the time parameters table togive it the current servicing time. The current servicing time is thenread and compared with the real time clock. The process

continually cycles around a loop, checking the real time against thecurrent servicing time for each algorithm, until the real time clockreaches the current servicing time for an algorithm.

When this occurs, the microprocessor exits the loop, reads the startaddress for the algorithm from the time parameters table andexecutes the algorithm.

After the algorithm has been executed, the microprocessor returnsto the loop as indicated by"return"in Figure 14.

The rotors in the generators of the heat pump may operate at highrotational speeds. For example the generators and heat pump maybe designed so that the rotors revolve at 15000 rpm or more. Tomaintain the performance of the generator at high revolutionspeeds, it is necessary to balance the rotating group (turbine, rotor,shaft and bearing system). Also, sealing the rotor and generatorinto the refrigerant cycle may avoid problems with losses andreliability of transferring power of the cycle through a shaft.

Furthermore, if a fixed magnet rotor is used, sensitive balancingbecomes difficult due to the magnetic field about the rotor and theferromagnetic

components of the rig become magnetised and if a sudden load isapplied to the generator, the resulting force can unbalance therotor.

The generator of the present invention includes a rotor that isnonmagnetic and can not become magnetised. The rotor may, forexample, be produced from Lycore 150 electrical sheet steel. The



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The total power output is computed as the product of IT and VT.This is compared to the previous power output. If the previouspower output was less than the current power output, the directcurrent through the stator windings is increased by a predeterminedstep size. If the previous power output was more than the currentpower output, the direct current through the stator windings isdecreased by a predetermined step size. Those skilled in the artwill appreciate that the algorithm illustrated in Figure 12 may useto control multiple target generators.

Where in the foregoing description, reference has been made tospecific components or integers of the invention having knownequivalents then such equivalents are herein incorporated as ifindividually set forth.

Although this invention has been described by way of example andwith reference to possible embodiments thereof, it is to be

understood that modifications or improvements may be madethereto without departing from the scope of the invention asdefined in the appended claims.

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