HEAT PUMP

This invention relates to a heat pump device particularly for air conditioning, refrigeration and heat pumping systems. The device relates especially to systems that have high energy efficiencies by recuperating heat or coolth that would otherwise be lost by conduction, radiation and convection.

In this specification the term 'heat pump' describes any powered device which moves heat from a source to a sink against a thermal gradient. A refrigerator is a particular type of heat pump where the lower temperature is required for the intended application. The term 'heat pump' is also used in a more limited sense than in this specification to describe a powered device which moves heat from a source to a sink against a thermal gradient where the higher temperature is required. The distinction between a refrigerator and a narrowly-defined heat pump is merely one of intended purpose, not operating principle. Indeed, many air conditioning systems are designed to supply either heating or cooling depending upon the user's need at a specific time.

In this specification the term "convection" includes any gas or liquid flow which may be the result of adventitious pressure or temperature differences or may be induced by a pump or fan. In this specification all pressures are given in "bars absolute", which is abbreviated to

"bara".

In this specification the energy efficiency of a device used for cooling is expressed by its "Coefficient of Performance" (COP (cooling)) which is the ratio of the energy supplied to drive the device to the quantity of energy removed in the cooling process. When a device is used for heating its COP (heating) is the ratio of the energy supplied to the quantity of heat produced in the heating process.

Certain aspects of this invention require the generation of large surface areas between gas and liquid phases using "static and/or jet mixers". A "static mixer" is an "in-line" device with no moving parts for efficiently blending (mixing) two fluid streams relying solely upon the energy in the fluid flows. A "static mixer" may consist of a tube containing mixer elements, which induce turbulence and shear in the flowing fluids and thus mixing them over the length of

the tube. The two fluids enter the mixer at one end as two separate phases and fluid blend exits from the other end of the tube. A "jet mixer" may comprise two feed channels that allow the two fluid flows to impinge at high velocity in a small volume causing considerable turbulence and thus creating a fluid blend which exits via a third channel. A mechanical mixer may be used. A mechanical mixer may comprise a moving element driven by a power source. The power source may be a motor which also drives the compressor or an expander actuated by a flow of working fluid. In the context of this specification a blend or mixture generated by a static, jet or mechanical mixer can be a solution, emulsion, aerosol or foam.

In describing the function of an air conditioning unit or refrigeration unit it is convenient to employ the concept of 'coolth' as the opposite of heat. If a substance is heated by the addition of 1 kJ then its heat content has been incremented by this amount. Conversely if IkJ of heat is removed from a substance then its coolth is said to have increased by this amount.

Although the primary purpose of an air conditioner is to maintain the temperature of a room at a chosen value, the comfort of occupants is also determined by humidity and the buildup of impurities. Periodic air changes are therefore also required. In older buildings such ventilation may occur by uncontrolled, adventitious leakage, i.e. outflow of room air balanced by the inflow of external air as draughts. In new buildings greatly improved sealing and better insulation requires that rooms be ventilated to remove stale air containing impurities such as volatile organic compounds from furnishings, odours and excess humidity. If warm fresh air is simply pumped in to balance the cool stale air expelled then energy supplied to the a/c unit will be required to cool the fresh air from the higher external temperature to the desired room temperature. This is an energy inefficient process because the coolth in the outgoing air is being wasted. By passing the cool outgoing air and the warm incoming air through a counter-current or cross flow heat exchanger the former can cool the latter. In other words at least part of the coolth of the outgoing air is transferred to the incoming air.

A process whereby heat is transferred from one fluid stream to another to reduce the load on a cooling unit is variously called 'coolth' recovery, regeneration or recuperation.

Conversely if a cold fresh air stream is being warmed by an exiting warm stale air stream then the process is called heat recovery, regeneration or recuperation. In the context of this specification heat or coolth recovery, regeneration and recuperation are considered to be equivalent. The heat exchanger used for this purpose can be described as a regenerator or a recuperator.

Coolth may be lost from a cooled space, or heat from a heated space, by air leaving and entering an enclosed space ("convection"), either adventitiously or because of forced ventilation, and/or through the enclosing walls predominantly by conduction or radiation. An important aspect of this invention is the integration of heat pumping with coolth or heat recovery or recuperation to minimise these coolth or heat losses and thus improve the overall energy efficiency of the total system.

In this specification the term "gas" is taken to include the term "vapour".

The term "temperature lift" refers to the difference in temperature between the heat sink and the heat source.

The term "working fluid" refers to a liquid or gas whose temperature can be changed, and optionally undergoes a change of state when an external force does work on it.

The term "heat transfer fluid" refers to a gas or liquid which absorbs or rejects heat or coolth to another material. In contrast to a "working fluid", a "heat transfer fluid" is not acted upon by an external force and the input of work in order to change its temperature.

In an especially preferred embodiment, the cyclical absorption and desorption of a volatile substance from a low volatility solvent is induced by a compressor.

In this specification the volatile substance and the solvent are collectively referred to as the working fluid. Both the volatile substance and the solvent can be single chemical compounds or mixtures.

The term "reactive gas" as used in this specification refers to a volatile substance which may be reversibly absorbed into a solvent.

In this specification, "lubricating oil", "oil" and "lubricant" are used as interchangeable terms.

An object of this invention is to provide means for the integration of heat pumping with the recovery of heat or coolth that would otherwise be lost by conduction, radiation and convection utilising working fluid, which, at constant pressure, boils over a temperature range.

A further object of this invention is to provide a working fluid composition that can be used in a high energy efficiency heat pumping device, which are intended for heat recovery, and in which the working fluid composition undergoes a partial or complete transformation between the liquid and vapour states at essentially constant pressure with progressive changes in temperature. For example, a liquid blend containing the hydrofluorocarbons Rl 34a and R32 in equimolar proportions at 5 bar will completely evaporate between -2.8 and 3.6 ⁰ C. A further example of working fluid is a solution containing 15% by weight of carbon dioxide in a polyester oil such as Emkarate™RH32L.

In the context of this specification the temperature glide of a wide glide working fluid is the difference between the temperature at which the working fluid enters the section of a heat exchanger where the fluid is undergoing a change of state and the temperature at which it leaves this section. In the context of this specification a temperature glide is considered to be "wide" if this temperature difference is 4 K or greater.

In this specification where a gas is in contact with a solvent, for example a lubricating oil, it is assumed that a portion of the gas will dissolve in the solvent to form a liquid solution whose concentration is determined by the temperature, pressure and rate of desorption or adsorption. Although the liquid may still be referred by the name of the solvent, or simply as an oil, it is to be understood that it may contain a quantity of dissolved gas.

According to the present invention, there is provided a heat pump comprising a circuit including: a first heat exchanger operatively connected to a heat sink; a second heat exchanger operatively connected to a heat source; a working fluid; pump or compressor; and an expansion device connected between the heat exchangers; the circuit being arranged so that working fluid is circulated between the heat exchangers by the pump or compressor so that working fluid passes successively from the pump or compressor to the first heat exchanger, the expansion device, the second heat exchanger and returns to the pump or compressor; the first heat exchanger including a first conduit for a heat transfer fluid; the second heat exchanger including a second conduit for a heat transfer fluid;

wherein the heat exchangers provide temperature glides to respective flows of heat transfer fluid with the temperature at one end of the first exchanger being approximately equal to the temperature of the heat transfer fluid leaving the heat sink; and the temperature at one end of the second heat exchanger being approximately equal to the temperature of heat transfer fluid leaving the heat source.

According to a first preferred aspect of this invention the working fluid comprises a higher volatility gas and a lower volatility liquid solvent capable of absorbing and desorbing the gas.

The volatility of the gas is higher than the volatility of the solvent.

In a preferred embodiment the working fluid also functions as a lubricant. Alternatively or in addition a separate lubricant may be employed and this may form a lubricant-rich liquid phase separate from a working fluid rich liquid phase,

This invention also relates to a heating or cooling device comprising a closed circuit wherein a heat transfer fluid, cooled or heated by a heat pump, is continuously or intermittently circulated in one direction through a heat exchanger which is partly or completely embedded in the material of a wall of an enclosed space which space is to be cooled or heated, to create and maintain a temperature gradient across the wall. This enables heat or coolth which may be lost by radiation or conduction through the wall to be recuperated.

Means for heat or coolth recuperation may be provided by the wall-embedded heat exchangers.

In a preferred embodiment of this invention there is provided means for the integration of heat pumping with coolth or heat recovery from one fluid stream and transferring the receiver heat or coolth to a second stream, which in turn transfers heat or coolth to or from a ground heat source.

In this specification, unless otherwise specified, the term "wall" is taken to include "side", "floor", "top" and "ceiling". An "embedded" heat exchanger is a heat exchanger that is located partially or wholly within or in physical contact with a wall.

In this specification the term "temperature glide" is defined as follows -

When a refrigerant comprises two or more components at least one component being a volatile material under the operating conditions of the heat pump, then at constant pressure the conversion of the volatile component or components from the gaseous to the liquid state may occur over a temperature range. Similarly when one or more volatile components are converted from the liquid to the gaseous state at constant pressure, this process may also occur over a temperature range. These ranges are called glides. The term glide gradients may be also used.

Although a blend of certain fluids with a specific composition may form an azeotrope at a specific temperature, at other temperatures this composition will show a glide. Blends of fluids that do not form azeotropes at any temperature are generally called zeotropes.

In preferred embodiments the glide may be in the range 5 to 100 K more preferably 10- 30K. The preferred glide depends upon the specific application.

All heat pumping equipment driven by an external energy source can potentially contribute to global warming through the combustion of fossil fuel which releases CO ₂ into the atmosphere. This is sometimes called 'indirect' global warming to distinguish it from 'direct' global warming caused by the release of refrigerants with high global warming potentials such as HFCs. In most heat pumps the indirect contribution significantly exceeds the direct effect, for example, by a factor of at least five. The total global warming caused by a heat pump can be reduced by making the device more energy efficient or by replacing the HFCs with refrigerants having lower global warming potentials. Preferably a combination of both approaches may be used.

In devices constructed according to this invention heat is supplied by and rejected to a heat transfer fluid located externally with respect to the device. The heat transfer fluid may be a gas, for example, air, or a liquid, for example water, ethylene glycol, propylene glycol or mixtures thereof, which do not undergo phase change transitions over the operating temperature ranges of the devices claimed in this specification. Alternatively, the heat transfer fluid might be a medium which undergoes at least a partial phase change during heating or cooling, for example, humid air. When fresh air is cooled below its dew point by the evaporator, water will condense out. In one embodiment of this invention condensed water is pumped into stale air flowing out from the room so that it will be vaporised, thus improving the energy efficiency of the system.

In a preferred embodiment there is provided a third heat exchanger with a conduit for heat transfer fluid connected in a circuit with either the first heat exchanger or the second heat exchanger and embedded in one or more walls of the enclosure to be heated or cooled; a means for circulating the heat transfer fluid between the third and either the first or the second heat exchangers.

The embedded third heat exchanger may be connected to the first heat exchanger when the enclosed space is to be heated or may be connected to the second heat exchanger when the enclosed space is being cooled.

Optionally the heat pump comprises a further heat exchanger to transfer heat from the working fluid in the line connecting the absorber to the expansion device, to the working fluid in the line suction connecting the desorber to the compressor.

In preferred embodiments, a mixer, for example, a jet mixer, static mixer or mechanical mixer may be provided. This may generate a large surface area between the gas and liquid phases to enhance the rate of adsorption. The mixer may cause the gas and liquid phases to form a foam, aerosol, emulsion or other highly mixed state in part or all of the circuit.

A surface active agent may be added to the working fluid. This may facilitate the generation of a large gas/liquid interface thus reducing the resistance to gas absorption.

A further heat exchanger may be located between the compressor and the high temperature heat exchanger. This may serve to remove heat of compression from the working fluid before the latter comes into contact with the first heat exchanger;

Optionally a gas-liquid separator may be provided between the desorber and the compressor; a liquid-gas mixer being provided between the compressor and the adsorber; and a liquid pump being provided between the separator and the mixer.

Optionally the cross-sectional area of the first temperature heat exchanger ("absorber") may progressively decrease in the direction of flow of the working fluid.

Optionally the cross-sectional area of the second temperature heat exchanger ("desorber") may progressively increase in the direction of flow of the working fluid.

A heat pump in accordance with this invention therefore may utilise the compressor or pump induced cyclical absorption and desorption of a volatile substance from a low volatility solvent.

The more volatile component of the working fluid may be a gas which can dissolve in a suitable solvent. The gas may be selected from the group consisting of: hydrocarbons, ammonia,

HFCs, fluorocarbon-iodides, unsaturated fluorinated compounds containing 2 to 6 carbon atoms, fluoro-olefins having the empirical formula C ₃ H _x F _6-X where x = 0 to 3, CO ₂ , and mixtures thereof.

The less volatile solvent component of the working fluid may be selected from the group consisting of water, saturated and unsaturated hydrocarbons, aromatic hydrocarbons, amines, alcohols, esters, ethers, amides, organic carbonates, hydro fluorocarbons and mixtures thereof.

Criteria for selecting a preferred combination of volatile component and solvent may include the following:

(i) Preferred solvents have normal boiling points at least 20 K higher than the boiling point of the volatile substance, more preferably at least 50 K higher and most preferably at least 100 K higher.

(ii) The solvent and the gas preferably have an affinity for each other. Preferably the solution of the gas in the solvent may show a positive deviation from Raoult's Law. Preferably the molar heat of solution of the gas in the solvent is at least lOkJ/Mole and more preferably should be greater than 15 kJ/mole.

(iii) The solvent should not react irreversibly with the gas.

(iv) The solvent when containing the lowest concentration of the reactive gas required in a device should remain liquid at the lowest operating temperature of the device.

(v) The viscosity of the solution in the system is sufficiently low so that it is driven forward through the heat exchangers by the gas flow.

The preferred volatile component is selected from the group consisting of carbon dioxide, HFC 134a, HFC 125, HFC 32, HFC 143a, HFC 227ea, HFC 245fa, CF ₃ CF=CF ₂ , CF ₃ CH=CF ₂ , cis- and trans- CF ₃ CF=CHF, cis- and trans CF ₃ CH=CFH and HFO 1234yf (CF ₃ CF=CH ₂ ).

The preferred solvent or lubricant may be a synthetic oil chosen from the group consisting of alkylbenzenes, polyvinyl ethers, polyalkylene glycols, amides, esters and mixtures thereof. Especially preferred are branched polyol esters having a molecular weight between 460 and 2000 and derived from polyols such as penta-erythritol and 1,1,1 -tris(hydroxymethyl)ethane esterifϊed with fatty acids having between 4 and 12 carbon atoms.

A further solvent may include propylene carbonate, ethylene carbonate or mixtures thereof. Preferably these solvents may be combined with a lubricant that forms a separate liquid phase. Especially preferred lubricants are hydrocarbons for example, mineral oils and alkyl benzenes.

Preferred volatile gases for this invention have Global Warming Potentials less than 150 and more preferably less than 10. In one preferred embodiment of this invention the reactive gas is CO ₂ . In another embodiment of this invention the reactive gas is a blend of HFC 125 and HFC 32 especially in a mass ratio of about 50:50. This has been designated R410A by ASHRAE. In another embodiment of this invention the reactive gas is HFO1234yf. In a further embodiment of this invention the reactive gas is HFO 1225ye (CF ₃ CH:CF ₂ ). In a further embodiment of this invention the volatile gas is a blend of CO ₂ and HFO 1234yf such that the blend when completely in the gaseous state is non-flammable as measured by ASTM E-681. In a further embodiment of this invention the volatile gas is a blend of CO ₂ and HFC 134a.

Preferred embodiments of this invention provide efficient transfer of volatile gas between the gas and solution phases. This is facilitated by the generation of a large surface area between the liquid and the gas. A large surface area can be achieved by incorporating static mixers, jet mixers, or mechanical mixers. Static mixers are available from a number of suppliers. One type is the Kenics mixer. Especially preferred is a combined static mixer and heat exchanger. Jet mixers also come in a variety of designs, for example the channels for the incoming fluid flows may meet at a T junction, a Y junction or a venturi nozzle. To promote absorption the mixers may be located in a high pressure part of the circuit. In one embodiment of this invention they are incorporated between the compressor and the absorber. In a further

embodiment one or more static mixers are incorporated within the absorber heat exchanger to ensure continued mixing.

Preferably the gas and liquid phases are mixed to form a foam in part or all of the circuit. This confers several advantages.

A mixer may be incorporated in low pressure part of the circuit between the expansion device and the desorber to enhance the rate of adiabatic desorption of the reactive gas from the oil. In another embodiment one or more mixers are incorporated in the desorber to enhance the gas desorption rate within the heat exchanger. In a further embodiment of this invention a mixer is incorporated between the desorber and the compressor to create a uniform foam which ensures that the mass ratio of gas to oil/absorbent sucked in by the compressor remains essentially constant without rapid fluctuations and also reduces the external sound level of the compressor.

Use of a foam promotes gas absorption on the high pressure side and promotes desorption on the low pressure side. Uniform distribution of the working fluid is facilitated in the absorber and desorber, especially when using multi channel exchangers e.g. flat plate heat exchangers. Suction into the compressor is also improved.

Surface-active agents or surfactants may be employed to generate a foam or an aerosol, in order to increase the contact surface area of the gas and the liquid and thus the rate of transfer of gas into the solvent in the absorber, and from the solvent in the desorber. Suitable surfactants may be monomeric, oligomeric or polymeric. Non-ionic surfactants are preferred for this invention. EP-A-I 252 278 discloses use of partially fluorinated ethers, notably Fomblin ™ HC- OH and Fluorolink™ ElO, when used with HFC- 134a and with air. Foams generated in this way can be used to reduce sound levels emitted by compressors. In the present specification fluorocarbon surfactants may be used with other gases such as CO ₂ and HFO 1234yf as well as HFCs to enhance absorption and desorption rates by increasing gas/solvent contact surface areas as well as to deaden sound emissions. Suitable concentrations may be in the range 10 ppm to 10000 ppm, preferably 50 to 1000 ppm and most preferably 100 to 500 ppm.

In a preferred embodiment all the oil or solvent together with the gas pass through the compressor so that the compressor serves to pump both components around the system.

Compressors suitable for this purpose must be capable of accepting a significant liquid flow, e.g. up to 30% by volume, without risk of physical damage. Suitable compressors include but are not

limited to moving vane, scroll, screw and liquid ring types. The latter three are preferred because they can accept substantial amounts of liquid flow. This liquid flow also provides good sealing to allow the bearings to be isolated from the displacement volume so that the solvent and the lubricant can be different.

In another preferred embodiment of this invention the lubricant and the absorbent are different liquids which are essentially immiscible over the operating conditions of the heat pump. The absorbent may be a good solvent for CO ₂ and may have a molar heat capacity which is preferably less than 60% that of the molar heat capacity of the lubricant, more preferably less than 40% and most preferably less than 20%. The lubricant may be a poor solvent for CO ₂ so that its viscosity may not be reduced by dilution with the gas.

The lubricant can thus be optimised for lubrication and the solvent for its ability to absorb the gas. For example mineral oil lubricant, which is a poor solvent for CO ₂ , can be used for the bearings and propylene carbonate, which is a good solvent for CO ₂ , can be used as the adsorbent. By good solvent is meant a liquid which maximises the mass of CO ₂ dissolved relative to the mass of the liquid at a specified pressure.

In one embodiment of this invention the non-miscible lubricant and the absorbent may be separated by collecting them in a vessel which allows them to settle into two liquid phases. In one aspect of this invention the vessel is located between the evaporator and the compressor suction port. Where the lower more dense liquid is the solvent, for example propylene carbonate, this may be drawn off and pumped from the lower pressure suction side of the compressor to the higher pressure gas discharge line of the compressor where it may be mixed with the higher pressure gas from the compressor discharge port. The upper, less dense oil, for example mineral oil, may be drawn off and allowed to drain to the compressor oil sump which provides an lubricant reservoir.

In another aspect of this invention the lubricant and liquid phases may be separated by causing them to rotate in a vessel, for example a cyclone or centrifuging device. Under the influence of the induced centripetal force the more dense liquid phase will tend to move the outer part of the vessel and the less dense liquid phase accumulate in the vicinity of the axis of rotation of the fluid.

In a further embodiment of this invention the immiscible lubricant and the solvent may be separated between the compressor discharge port and the entry to the higher temperature heat exchanger ("absorber"). In this case the lubricant is driven by the pressure difference between compressor suction and the discharge ports back to the compressor sump.

The maximum operating pressure may be in the range 8 to 35 bar. For equipment constructed from components employed in HFC units the preferred the maximum operating pressure is preferably in the range 8 to 30 bar more preferably in the range 8 to 25 bar.

In a further embodiment of this invention a heat pump may be employed to cool the condenser of a Rankine cycle refrigeration unit containing CO ₂ as the working fluid. The evaporating temperature of the CO ₂ in the Rankine cycle unit may be in the range -35 to 5 ⁰ C and may be used to provide cooling for example to keep food frozen or to generate ice. In another embodiment of this invention a secondary refrigerant, for example water, may be used to transfer coolth from the desorber of a heat pump to the condenser of a CO ₂ Rankine cycle unit.

According to a second preferred aspect of the present invention the working fluid in the heat exchangers may have a wide glide.

The glide may be determined by measurement of the temperature profile of the evaporator or condenser, for example using a thermocouple.

A wide glide is a temperature difference of about 4K or greater, for example about 8 to about 12K.

A wide temperature glide working fluid will necessarily comprise at least two component single fluids. The fluids and their relative proportions may be selected so that the thermodynamic properties of the blend deliver the performance parameters required from a heat pump in its intended application. Various factors may be taken into account in formulating a wide temperature glide working fluid.

The temperature of a heat transfer fluid, which does not under go a phase change, progressively increases or decreases as it travels along a heat exchanger. The composition of the working fluid is selected such that its evaporating or condensing temperature also progressively increases or decreases respectively as it travels along a heat exchanger counter-current to the

heat transfer fluid. For high energy efficiency, the temperature profile of the working fluid may be closely matched to the temperature profile of the heat transfer fluid. The temperature profiles of the heat transfer fluid and the working fluid necessarily differ by a finite temperature difference to provide a temperature gradient across the walls of the heat exchanger, which drives heat transfer. This temperature difference is sometimes called the "approach temperature". In case of an evaporator or desorber, the temperature of the working fluid will be lower than that of the heat transfer fluid at any point along the evaporator. In the case of a condenser or absorber the temperature of the working fluid will be higher than that of the heat transfer fluid at any point. To maximise further energy efficiency, an approach temperature at any point in a heat exchanger is preferably as small as possible. However the smaller the approach temperature the larger the heat exchanger surface area required for the transfer of given quantity of heat, which increases the heat exchanger cost. Approach temperatures in the range of 1 to 10 K and preferably 2 to 5 K provide an acceptable compromise between energy efficiency and cost. The composition of the working fluid blend is chosen such that the approach temperature along a heat exchanger is essentially constant, where a change in state of the working fluid is occurring.

The temperature glide required for a working fluid may be determined by the application for which it is intended. The following examples are illustrative, but not limitative. For ground source heat pumps, a temperature glide in the range about 5 to about 12 K is preferred. For air to air reversible heat pumps a glide in the range about 8 to about 25 K may be preferred. For refrigeration equipment a glide of about 15 to about 35 K may be preferred. For biomedical refrigeration applications glides in the range about 20 to about 100 K may be preferred.

Selecting a working fluid composition with a required temperature glide is not the sole criterion in selecting the components of a working fluid. The maximum operating pressure should not exceed the pressure rating of the heat pumping device. The composition of the fluid blend may be selected such that its pressure in the condenser may not exceed 1.5 times the test pressure of the equipment. The maximum operating pressure of a heat pump may be determined by the maximum condenser or absorber temperature at which it is required to run. It is disadvantageous to select a working fluid composition which has too low a vapour pressure. A low evaporating pressure produces a low suction density, requiring a larger compressor with a higher swept volume. Such a compressor may be more expensive and may also occupy more space, and is disadvantageous where price and compactness are important, for example in small a/c units designed for domestic use. Working fluid compositions should therefore be chosen to

maximise the evaporating or desorbing pressures without exceeding the maximum pressure rating of the equipment.

Component zeotropic fluids which are suitable wide glide working fluids may be commercially available. These meet ASHRAE Al, A2, Bl or B2 safety classifications and have Global Warming Potentials (GWP) that do not exceed 5000. Preferred fluids may be CO ₂ , R32, R125, R143a, R134a, R227ea and R245fa and mixtures thereof.

Component fluids with GWPs equal to or less than 150 are especially preferred. These include but are not limited to CO ₂ , R152a and R1234yf and mixtures thereof.

For applications where natural refrigerants are required preferred component fluids are selected from the group comprising: CO ₂ , ethane, propane, isobutane, butane, neopentane, isopentane and pentane and mixtures thereof.

Table 1 provides examples of suitable zeotropic blends.

The components of working fluid may be selected from the group consisting of: acylic and cyclic saturated and unsaturated hydrocarbons containing 1 to 6 carbon atoms; saturated acyclic and cyclic hydrofluorocarbons (HFCs) containing 1 to 6 carbon atoms; acyclic and cyclic unsaturated hydrofluorocarbons having 2 to 6 carbon atoms; fluorocarbon-iodides; saturated fluorocarbon ethers having 2 to 6 carbon atoms; CO ₂ ; and mixtures thereof.

Suitable hydrocarbon components include methane, ethane, propane, isobutane, butane, normal pentane, 2-methylpentane, 3-methylpentane, 2,2-dimethylpropane, cyclopropane, cyclo- butane, cyclo-pentane, ethane, propene, 1-butene, and 2-methylpropene and mixtures thereof.

Suitable hydrofluorocarbons include, but are not limited to, acyclic alkanes represented by the formula C _n H( _2n+2-X )F _x where n = 1 to 6 and 0<x<2n+2 and preferably x>3n+l-x for non- flammability; acyclic alkenes represented by the formula C _n H _(2n _ _X) F _x where n = 1 to 6 and x >0 and preferably x>3n-x for nonflammability; saturated ethers represented by the formula C _n H( _2n+2-X )F _x O where n = 1 to 6 and 0<x<2n+2 and preferably x>3n-x for non-flammability.

Component fluorocarbon fluids suitable for a wide glide blend required by this invention include, but are not limited to, CH ₂ F ₂ , CF ₃ CF ₂ H, CF ₃ CH ₃ , CF ₃ CH ₂ F, CF ₃ CHFCF ₃ , CF ₃ CF ₂ CH ₃ , CF ₃ CH ₂ CF ₃ , CHF ₂ CF ₂ CH ₂ F, CF ₃ CH=CF ₂ , cis-CF ₃ CF=CFH, trans-CF ₃ CF=CFH,

CF ₃ CF=CH ₂ , CiS-CF ₃ CH=CHF, trans-CF ₃ CH=CHF, CF ₃ OCH ₃ , CF ₃ CHFOCHF ₂ , CHF ₂ CF ₂ OCH ₃ , CF ₃ CF ₂ CF ₂ CHF ₂ , CF ₃ CF ₂ CF=CH ₂ , and 0(-CF ₂ CF ₂ CH ₂ CH ₂ -).

Carbon dioxide may be employed.

Criteria for selecting the components for a wide glide zeotropic blend include, but are not limited to, the following properties.

The components should not react irreversibly with themselves or with other materials in the heat pump under operating conditions.

The maximum operating pressure of the blend should not exceed 30 bara and preferably it may not exceed 25 bara. More preferably the maximum operating pressure may not exceed 20 bara.

Preferably the blend may be rated class 2 according to the ASHRAE 34 flammability standard. More preferably it may be rated class 1. For an application where use of a flammable refrigerant may be acceptable, hydrocarbons blends may be especially preferred. For an application where low flammability or non-flammability may be required, fluorine-containing, alkanes, alkenes and ethers may be preferred. For a component to be rated class 1, i.e. nonflammable according to the ASHRAE 34 Standard, the number of carbon-carbon bonds plus the number of carbon- hydrogen bonds should be less than the number of carbon-fluorine bonds. A blend may contain one or more components which are flammable provided the blend itself is not flammable according the criteria laid out in the ASHRAE 34 Standard.

The blend of refrigerant fluids may be selected to provide both a temperature glide similar in value to the temperature difference between the entry and exit temperatures of the heat transfer fluid passing through the evaporator, and an operating pressure that falls within the desired range.

For example in a typical room air conditioning system a difference between the entry and dew point of a working fluid may be in the range at least 10 to 20 K with an operating pressure of around 4 to 10 bara.

If a heat pump is being used for heating then component refrigerant fluids comprising a blend may be selected to combine a temperature glide similar in value to the temperature difference between the entry and exit temperatures of the heat transfer fluid passing through the

condenser, and an operating pressure which falls within the desired range. In the evaporator the temperature glide may also include a further contribution from superheating the vapour above the dew point to minimise the risk of liquid refrigerant returning to the compressor. In the condenser the temperature glide may also contain a further contribution from desuperheating the vapour above its dew point and from subcooling the liquid refrigerant below its bubble point. The contribution may be about 5 K.

Lubricants which have been used with HFC and hydrocarbon lubricants may be used with the wide glide refrigerant blends employed in this invention. These lubricants include, but are not limited to, types commonly described as mineral oils, poly-alpha olefins, alkylbenzenes, polyalkylene glycols and polyol esters. If a refrigerant blend comprises HFCs, preferred lubricants are polyalkylene glycols and polyol esters. For non-fluorine containing refrigerant blends lubricants containing only carbon and hydrogen, mineral oils, poly-alpha olefins, and alkylbenzene lubricants are preferred. Non-flammable or low flammability hydrofluorocarbon blends may contain 0.5% to 5% (w/w) of a single hydrocarbon or a hydrocarbon mixture to assist the return of mineral oils, poly-alpha olefins, alkylbenzenes lubricants from the heat pump circuit to the compressor.

The present invention may be employed in a variety of applications, particularly those requiring the efficient pumping of heat or coolth between a higher temperature and a lower temperature. The following examples serve to illustrate the invention.

In one embodiment of this invention a heat pumping device is used to air condition a room. Although the primary purpose of an air conditioner (a/c) is to maintain the temperature of a room at a chosen value, the comfort of occupants is also determined by humidity and the build- up of impurities. Periodic air changes are therefore also required. In an older building such ventilation may occur by uncontrolled, adventitious convection, i.e. outflow of room air being balanced by the inflow of external air as draughts. In a new building greatly improved sealing and better insulation may require that a room be ventilated to remove stale air containing impurities, for example volatile organic compounds from furnishings, odours and excess humidity. If warm fresh air is simply pumped in order to balance the cool stale air expelled then energy supplied to the a/c unit will be required to cool the fresh air from the higher external temperature to the desired room temperature. This is an energy inefficient process because the coolth in the outgoing air is being wasted. By passing the cool, outgoing air and the warm incoming air through a counter-current or cross flow heat exchanger the former can cool the

latter. In this way a proportion of the coolth of the outgoing air is transferred to the incoming air. But the requirement for a finite temperature difference between the outgoing and incoming air at each point in the heat exchanger to drive heat transfer, plus heat gained by the air in the room, for example from internal sources such as computer equipment and by conduction through walls from outside, means that the outgoing air cannot cool the incoming air to the required low entry temperature. A further disadvantage of this simple heat recovery system is that the air exit and entry ports are constrained to be in relatively close proximity so that a part of the stale air expelled from a building may be sucked back in with fresh air.

Heat pumps in accordance with this invention, may incorporate features of one or both preferred aspects as set out above, that is by use of gas/solvent systems with wide glides.

The invention is further described by means of example but not in any limitative sense with reference to the accompanying drawings, of which: Figure 1 is a diagrammatic representation of a ventilation system in accordance with the invention;

Figure 2 is a diagrammatic representation of a further embodiment of the invention; Figure 3 is a diagrammatic representation of a further embodiment of the invention; Figure 4 is a diagrammatic representation of a further embodiment of the invention, including an underfloor water coil;

Figure 5 is a diagrammatic representation of a further embodiment of the invention, including ceiling and wall mounted heat exchangers;

Figure 6 is a diagrammatic representation of a reversible heat pump in accordance with this invention; Figure 7 is a diagrammatic representation of a further embodiment of the invention;

Figure 8 is a diagrammatic representation of a split heat pump in accordance with the invention;

Figure 9 is a schematic view of a further embodiment of the invention; Figure 10 is a schematic view of a further embodiment of the invention; Figure 11 is a schematic view of a further embodiment of the invention;

Figure 12 is a schematic view of a further embodiment of the invention; Figure 13 is a schematic view of a further embodiment of the invention Figure 14 is a schematic view of a further embodiment of the invention; Figure 15 is a schematic view of a further embodiment of the invention; Figure 16 shows an embedded heat exchanger section;

Figure 17 is a cross section through a heat exchanger in accordance with the invention; and

Figure 18 is a schematic view of a further embodiment of the invention.

Figure 1 shows a ventilation system designed to ventilate and air condition a room (1.1) comprising an absorbent heat pump to recuperate coolth from outgoing stale air and transfer it to incoming fresh air. The purpose of the device is to provide a combination of ventilation and air conditioning to maintain the temperature of a room at 5 to 20 K below the external temperature accompanied by the recuperation of coolth from the outgoing air and transfer of it to the incoming fresh air to provide good energy efficiency.

The device comprises a scroll compressor (1.2) which is capable of simultaneously compressing CO ₂ gas and pumping a polyol ester (POE) lubricant. The lubricant also acts as an absorbent for the CO ₂ . An adsorber (1.3) in which the CO ₂ dissolves in the POE is driven by pressure and the transfer of heat. An expansion valve (1.6) is provided; a desorber (1.7) allows the desorption of CO ₂ from the POE solution driven by the reduced pressure and the imput of heat. A blower (1.5) sucks stale air from the room and blows over the adsorber (1.3). Blower (1.9) sucks outside fresh air over the desorber (1.7) and blows it into the room. Control instrumentation monitors the operation of device and adjusts the expansion valve. The combination of CO ₂ and POE comprises the working fluid.

The room ( 1.1 ) is entered by a door (1.10) and is lit by a window (1.11).

Figure 1 shows a cross-section through the room with a ventilating system in accordance with this invention. Scroll compressor (1.2) sucks in a mixture of CO ₂ and a weak solution of CO ₂ in a polyol ester Emkarate RL32H (POE) that acts as both absorbent and lubricant. The mixture is compressed and after discharge from the compressor passes through the in-line static mixer (1.12).

A portion of the gas dissolves in the POE.

The temperature rises because of the compression and the latent heat of solution generated by the gas. The compressed gas/solution is discharged into the absorber (1.3) where it is progressively cooled by transferring heat to the stale air stream (1.4). The stale air is sucked from the room by fan (1.5). As the CO ₂ gas/solution travels along absorber (1.3) the gas

progressively dissolves in the liquid until it reaches the expansion valve (1.6) where the gas is mainly or wholly in solution. The solution plus any residual gas travels through the expansion valve where adiabatic flashing of the CO ₂ from the solution commences and cools the oil/gas mixture. The mixture enters the desorber (1.7) where it is progressively heated by the fresh airflow (1.8) being blown into the room by fan (1.9), while the air flow becomes progressively cooled. The increasing temperature of the solution in the desorber causes more CO ₂ to desorb. After exiting the desorber the gas/oil mixture enters the suction port of the compressor to complete the circuit. The system shown in Figure 1 is intended to both ventilate the room with at least four air changes per hour and to maintain the room at 20 ⁰ C with an external temperature of up to 40 ⁰ C.

In another embodiment of this invention R410A is used in place of CO ₂ in the device represented in Figure 1.

A further embodiment of this invention is represented by the heat pumping device shown in Figure 2 where it is employed to heat a room (2.1) which can be entered by door (2.10) and is lit by window (2.11). Figure 2 shows a cross-section through the room. The device contains a mixture of CO ₂ and polyol ester oil Emkarate™ RL32H (POE) that acts as both absorbent and lubricant. Scroll compressor (2.2) sucks in CO ₂ from gas separator (2.12) together with sufficient oil for lubricating the bearings and rotor. The major part of the oil flow is pumped from separator (2.12) by the liquid pump (2.13) into jet mixer (2.14) where it meets the compressed gas from compressor (2.2). The jet mixer comprises channels in the form of letter "T" whose diameters are smaller than the diameters of the gas and liquid feed pipes and the exit pipe for the gas/liquid mixture. The pressure drop across the mixer is sufficient to create a large mixing power input per unit volume at the point where the arms of the "T" meet. In this embodiment the liquid is fed into stem of the "T" and the gas into one side of the crossbar. The oil/gas mixture exits via the other side of the crossbar. To enhance the gas/liquid interface Fomblin™ fluorinated non-ionic surfactant may be dissolved in the oil at a concentration of 250 ppm.

The compressed gas/solution in the form of a foam is discharged into the absorber (2.7) where it is progressively cooled by transferring heat to the fresh air stream (2.8) pushed into the room by fan (2.9). As the CO ₂ solution travels along the absorber (2.7) the gas progressively dissolves in the liquid until it reaches the expansion valve (2.6) where the gas is mainly or wholly in solution. The solution plus any residual gas travels through the expansion valve where

adiabatic flashing of the CO ₂ from the solution commences and cools the oil/gas mixture. The mixture enters the adsorber (2.3) where it is progressively heated by the stale airflow (2.4) being sucked from the room by fan (2.5), while the airflow becomes progressively cooled. The increasing temperature of the solution in the desorber causes more CO ₂ to desorb. After exiting the desorber, the gas/oil mixture enters the gas/oil separator (2.12) to complete the circuit. The system shown in Figure 2 is intended to both ventilate the room with at least 4 air changes per hour and to maintain the room at 20 ⁰ C with an external temperature down to -20 ⁰ C.

Heat, known as geothermal energy, generated deep in the Earth by nuclear fission, flows outwards to the planet's surface where it is lost by radiation into space. In a further embodiment of this invention a heat pump containing a wide glide refrigerant blend may be employed to intercept geothermal energy and pump it to a higher temperature. Figure 3 represents a heat pump which uses geothermal energy as a heat source to heat a room (3.1) which can be entered by a door (3.10) and is lit by a window (3.1 1). A dotted line (3.8) indicates the surface of the ground.

Compressor (3.2) sucks in a two-phase mixture of lubricant/absorbent Emkarate™ RH32L and CO ₂ gas and compresses it into a heat exchanger channel (3.15) which transfers sensible heat from the fluids into water flowing through channel (3.16) of the same heat exchanger. The two-phase mixture then passes through an in-line static Kenics™ mixer (3.17) to produce a foam-like composition with greatly increased surface area between the gas and the oil at the exit of the mixer compared to the beginning of the cycle. The large surface allows adiabatic absorption of the CO ₂ in the lubricant in the mixer. The foam-like composition enters the channel (3.7) of a second heat exchange where it progressively loses heat to the counter- current water flow passing through channel (3.5) of the same heat exchanger. The reduction in temperature of the two-phase fluid results in more gas dissolving in the liquid. The fluid mixture then passes through channel (3.19) of a heat exchanger where heat is transferred to cooler low pressure CO ₂ /oil mixture flowing through channel (3.18) of the same heat exchanger. When the high pressure mixture emerges from (3.19) essentially all the gas has been absorbed in the oil and the solution passes through expansion valve (3.4) where its pressure is rapidly dropped causing the rapid adiabatic desorption of CO ₂ . This resulting cold gas/oil two-phase flow passes through heat exchanger channel (3.3) where it progressively gains heat from the counter-current current of water/glycol flowing through channel (3.14) of the same heat exchanger. As the temperature rises more CO ₂ is desorbed from the oil. After exiting from channel (3.3) the two- phase gas/oil flow passes through channel (3.18) where it gains more heat resulting in the further

desorption of CO ₂ . The low pressure gas/oil two-phase flow then re-enters the compressor suction port thus completing the cycle. The water pumped through heat exchanger channels (3.5) and (3.16) by pump (3.6) passes through under-floor heat exchanger coil (3.9) where it loses heat to room (3.1). The cooled water/glycol returns to heat exchanger 3.5. The water/glycol mixture cooled by the desorber (3.3) and pumped by pump (3.13) enters heat exchanger coil (3.12). The coil 93.12) is buried at least 1 metre below ground level (3.8), where it is progressively warmed by heat transfer from the surrounding soil before re-entering heat exchanger channel (3.14). The device shown in Figure 3 thus extracts low temperature from the ground, raises it a higher temperature which can then be used to heat a room (3.1). The soil, cooled by extraction of heat, regains its original temperature by heat flowing from the Earth's interior (geothermal heat) and by heat absorbed from solar radiation at the ground surface. The ground coil may be 120 metres long and can either be arranged laterally below the ground surface, as indicated in Figure 3, or where external area is limited sunk into a pit 50 to 100m deep. The system shown in Figure 3 is intended to maintain the room at 20 C with an external temperature of down to -20 C.

In, a further embodiment of this invention a heat source for a heat pump which heats room by an underfloor water circuit may be a solar panel. Radiant energy from the sun at visible and near infra-red wavelengths is absorbed by the roofs and walls of buildings and then reradiated at longer wavelengths and thus lost. A solar panel may capture the incoming radiation and use it to heat a liquid flowing through tubes in the panel. The warming liquid may be pumped through the evaporator of a heat pump which extracts heat from liquid thus cooling it. The cold liquid returns to the solar panel.

In another embodiment of this invention shown in Figure 4 where the heat source is a room (4.1) which is being heated by an under-floor water coil (4.12). The heat pump is similar in design and operation to that shown in Figure 3. In this embodiment the water/glycol circuit additionally includes heat exchangers (4.20) and (4.21) embedded in the ceiling and walls of the room respectively for recovering or recuperating heat escaping through the insulation of the room. This has the advantage of making the heating system more energy efficient and reducing the size and thus cost of the ground coil heat exchanger (4.12). The system shown in Figure 3 is intended to maintain the room at 20 ⁰ C with an external temperature of down to -20 ⁰ C.

In a further embodiment of this invention shown in Figure 5 a heat pump device, which is similar in design and operation to that shown in Figure 3, circulates cold water through ceiling and wall mounted heat exchangers (5.9) and (5.20) respectively which thus cool a room (5.1).

The heat is rejected via the heat exchanger channel (5.15) and adsorber (5.7) to water/glycol circulating through heat exchanger channels (5.16) and (5.5). The water/glycol is then cooled through the ground coil (5.12). The ground in the vicinity of the coil is thus warmed. The device shown in Figure 5 can be described as a ground sink heat pump or air conditioning chiller. The special advantage of this design over conventional air cooled air conditioners is the fact that the ground temperature below ~ 1 metre depth is typically 10 to 12 ⁰ C, which is 10 to 25 K cooler than the external air temperature thus providing high energy efficiency. The system shown in Figure 5 is intended to maintain the room at 20 ⁰ C with an external temperature up to 40 ⁰ C.

In a further embodiment of this invention the device includes both under-floor heating coils and ceiling/wall heat exchangers. The heat pump is designed to be "switchable" so that when air conditioning is required it delivers cold water/glycol to fanned ceiling/wall heat exchangers in the room, or conversely when heating is required it delivers warm water to these heat exchangers. During the day when air conditioning is required the heat pump rejects heat to the ground, while at night when heating might be required the heat pump recovers the rejected heat from the ground. This recuperation process enhances the energy efficiency of the device. This embodiment, called a reversible heat pump, is shown in Figures 6a and 6b where a fanned heat exchanger is used either to heat or air-condition the room (6.1) parts of whose floor and walls are represented by (6.10) and (6.11) respectively. For clarity the whole room is not shown. The system comprises a scroll compressor (6.2) through which the all the oil and the CO ₂ are pumped; a heat exchanger comprising channels (6.15) and (6.16); a heat exchanger comprising channels (6.5) and (6.7); a heat exchanger comprising channels (6.18) and (6.19); a heat exchanger comprising channels (6.3) and (6.14); room heat exchanger (6.9); ground coil heat exchanger (6.12); expansion valve (6.4); an in-line Kenics™ static mixer (6.17); water glycol pumps (6.6) and (6.14); and three-way valves (6.10), (6.11), (6.20) and (6.21).

Figure 6a shows valves (6.10, 6.11, 6.20 and 6.21) set in positions to enable the room (6.1) to be heated. Compressor (6.2) sucks a mixture of CO ₂ gas and oil/absorbent which contains a Fomblin™ fluorochemical surfactant to induce foaming. The hot, high pressure, gas/oil mixture discharged from the compressor enters the heat exchanger channel (6.15) where it rejects heat to water/glycol in channel (6.16) being pumped by (6.6). The gas/oil mixture then enters mixer (6.17) where the combined effect of the mixer and the surfactant great increases the surface area between the oil and the gas thus enhancing adiabatic adsorption. After leaving (6.17) the gas/oil mixture is further cooled by passing through channel (6.7) where it rejects heat to the water/glycol in channel (6.5). More gas is progressively adsorbed as the temperature of the

gas/oil falls and continues through heat exchanger channel (6.19) where more heat is rejected to the low pressure gas/oil mixture passing through heat exchanger channel (6.18). When the high pressure mixture emerges from (6.19) essentially all the gas has been absorbed in the oil and the solution passes through expansion valve (6.4) where its pressure is rapidly dropped causing the rapid adiabatic desorption Of CO ₂ . This resulting cold gas/oil two-phase flow passes through heat exchanger channel (6.3) where it progressively gains heat from the counter-current current of water/glycol flowing through channel (6.14) of the same heat exchanger. As the temperature rises more CO ₂ is desorbed from the oil. After exiting from (6.3) the two-phase gas/oil flow passes through (6.18) where it gains more heat resulting in the further desorption of CO ₂ . The low pressure gas/oil two-phase flow then re-enters the compressor suction port thus completing the cycle. The water pumped through heat exchanger channels (6.5) and (6.16) by pump (6.6) passes through room heat exchanger coil (6.9) where it loses heat to room (3.1). The cooled water/glycol returns to heat exchanger (6.5). The water/glycol mixture is cooled by the desorber (6.3) and pumped by pump (6.13) so that it enters heat exchanger coil (6.12) which is buried at least 1 metre below ground level (6.8). The mixture is progressively warmed by heat transfer from the surrounding soil before re-entering heat exchanger channel (6.14). The device shown in Figure 6a thus extracts low temperature heat from the ground, raises it a higher temperature which can then be used to heat room (6.1). The soil, cooled by extraction of heat, regains its original temperature by heat flowing from the Earth's interior (geothermal heat), by heat absorbed from solar radiation at the ground surface, or by heat rejected during a previous air conditioning operation. The ground coil may be 120 metres long and can either be arranged laterally below the ground surface, as indicated in Figure 6, or where external area is limited sunk into a pit 50 to 100m deep.

Figure 6b represents the device with the valves set to the position for the device to cool a room (6.1). The heat pump operates as described previously for heating room (6.1) (Figure 6a). For cooling the room the cool water/glycol from heat exchanger channel (6.14) is pumped through fanned heat exchanger (6.9) by pump (6.13), while a warm water/glycol mixture from heat exchanger (6.5) is pumped through ground coil (6.12) by pump (6.6). In this mode of operation heat is transferred from room (6.1) to ground from where it may be extracted when the heat pump operates in heating mode.

The system shown in Figure 6 is intended to maintain a room at 20 ⁰ C with an external temperature in the range -20 ⁰ C to 40 ⁰ C.

In another embodiment of this invention R410A is used in place of CO ₂ in the device represented in Figure 6.

In another embodiment of this invention the heat pumping device shown in Figure 6 has the absorber directly heated, in the case of room heating, or cooled, in the case of room cooling, by an air stream driven by a fan.

At present the vehicle air conditioning systems, especially those in automobiles, employ HFC 134a as the refrigerant in a standard Rankine Cycle device. Since HFC 134a has a higher global Warming Potential (1300 compared to 1 for CO ₂ ) The European Union has enacted legislation for its phase out starting with new road vehicle models from 1 ^st January 2011. Replacement refrigerants for Rl 34a must have a GWP not greater than 150. To ensure that the reduction in global warming achieved by eliminating HFC 134a is not negated by greater fuel consumption any new mobile air conditioning (MAC) systems must not be less energy efficient than the 134a systems they are replacing. Additionally automobile manufacturers prefer refrigerants that are non-flammable; have acceptably low toxicity; will fit into the limited space available for the MAC system; do not add extra weight to the vehicle; and operate at pressures comparable with 134a so are similar in design to existing systems.

Transcritical CO ₂ systems provide an acceptable combination of low toxicity, non- flammability and small size. However operating pressures are as high as 150 bar which is ~ 6 times higher than existing HFC 134a systems so that designs must differ considerably from HFC 134a units. Furthermore good energy efficiency has proven difficult to achieve because the high energy of compression imparted to the gas that needs to be recovered.

When a refrigerant blend compromising two refrigerants with widely differing vapour pressures such as CO ₂ with HFO 1234yf or Rl 34a is employed in a refrigerating/heat pumping system the CO ₂ accumulates in the condenser predominantly as gas causing generate a pressure which exceeds the maximum operating pressure of system. Furthermore, the systems claimed are typically designed so that the major part of the lubricant remains in the compressor at all times with only a minor proportion of the total, typically 2% by weight or less based on the circulating refrigerant passing around the circuit.

A further embodiment of this invention overcomes the separate limitations of CO ₂ and HFO 1234yf by employing a mixture of these two volatile substances absorbed in a lubricant as

the working fluid in an MAC system. In this embodiment sufficient amount of lubricant is circulated around the circuit to act as an absorbent ensuring that the combined pressures of the volatile components is not less than about 3 bara and preferably not less than about 3.5 bara at 5 ⁰ C in the desorber and preferably does not exceed about 25 bara and preferably not more than about 22 bara at 70 ⁰ C in the absorber.

A MAC system employing CO ₂ and HFO 1234yf absorbed in a lubricant is exemplified with reference to Figure 7.

The system comprises a scroll compressor (7.2); heat exchanger (7.3); in-line mixer

(7.4); absorber (7.5); fan (7.6); heat exchanger channel (7.7); expansion valve (7.8); desorber (7.9); fan (7.10); and heat exchanger channel (7.1 1). Compressor (7.1) sucks in a mixture of POE lubricant and gas containing CO ₂ and HFO 1234yf. After discharge from the compressor the mixture passes through heat exchanger (7.3) where it is cooled by and air stream driven by fan (7.6). The mixture then enters the in-line mixer (7.4) where the gas and liquid phases are thoroughly mixed to enhance the rate of adiabatic adsorption of the gases in the oil. The mixture then passes to heat exchanger (7.5) where it is further cooled by the air stream driven by fan (7.6) resulting in the progressive absorption of the gases in the oil. The mixture then enters heat exchanger channel (7.7) where further cooling occurs by heat transfer to the low pressure gas/oil mixture passing counter current through heat exchanger channel (7.11). When the working fluid mixture leaves (7.7) essentially all the gas will be dissolved in the oil and enters the expansion device (7.8) where the rapid drop in pressure causes the CO ₂ and HFO 1234yf to flash evaporate generating a cold mixture of gas and oil. The mixture enters heat exchanger (7.9) where it is heated by extracting heat from the air stream driven by fan (7.10). More CO ₂ and HFO 1234yf progressively desorb from the oil temperature of the mixture rises. The cooled air then enters to the passenger compartment (7.1) of the vehicle. The low pressure gas/oil mixture then passes through heat exchanger channel (7.1 1) where it is further heated by heat transfer from the high pressure refrigerant fluid in channel (7.7) before returning the compressor suction port to complete the cycle.

The preferred compositions of the CO ₂ /HFO 1234yf/lubricant comprising the working fluid are selected so that its pressure is in the range 5 bar to 9 bar absolute (bara) and more preferably in the range 6 to 8 bara, and should a leak occur in the air conditioning system the composition of the gas mixture that is ejected from the air conditioning system will be non- flammable as determined by ASTM E-681. More preferred compositions of the working fluid

meet the ASHRAE Committee 34 rating "1" for flammability. In especially preferred compositions the ratio of HFO 1234yf and CO ₂ in the absence meet the ASHRAE Committee 34 rating "1" for flammability.

In a further embodiment of this invention the system shown in Figure 7 is simplified by the omission of the heat exchanger comprising channels (7.7) and (7.11).

In a further embodiment of this invention the system shown in Figure 7 is simplified by the omission of the in-line mixer.

In a further embodiment of this invention the working fluid is a combination of a CO ₂ /134a/POE lubricant

In a further embodiment of this invention the working fluid of the system shown in Figure 7 also comprises a non-ionic fluorinated surfactant that causes the oil to foam.

In a further embodiment of this invention the working fluid is a combination of HFO 1225 ze, HFO 1234yf, CO ₂ and a POE oil. This composition is especially preferred because the non-flammable HFO 1225yez and CO ₂ have boiling points that are respectively higher and lower than that of flammable HFO 1234yf. Should a leak occur in a heat pump or storage vessel containing this working fluid the initially escaping gas will have a higher concentration of CO ₂ than the formulated composition and thus will suppress the flammability of HFO 1234yf. As the leak continues the concentration of CO ₂ will fall but the concentration of HFO 1225ze will increase and thus make a greater contribution to suppressing the flammability of HFO 1234yf. This combination of two non-flammable and one non-flammable volatile fluids, which can be described as a "non-flammable sandwich", ensures that the working fluid does not generate a flammable composition throughout the course of a leak from a heat pump or storage vessel.

In further embodiments of this invention the working fluids are blends of HFO 1225yez (Z-CF ₃ CH=CF ₂ ) / CO ₂ with a lubricant, especially a POE. This combination is especially preferred because both and CO ₂ and HFO 1225yez are non-flammable. However HFO 1225yez boils at -19 ⁰ C so has a significantly lower capacity than HFC 134a, the refrigerant presently used in MAC. The addition of CO ₂ increases the total pressure of the volatile components and hence the capacity of the working fluid. By employing the POE both as a lubricant and as well

as the absorbent for the volatile more CO ₂ can be added thus reducing the GWP of the blend and its cost relative to pure HFO 1225yez.

A further embodiment of this invention is the split air conditioning unit shown in Figure 8. This it is employed to cool a room (8.1) which can be entered by door (8.10) and is lit by skylight (8.11). The device contains a mixture of CO ₂ , propylene carbonate as solvent, and ISO grade 32 alkylbenzene as lubricant. Scroll compressor (8.2) sucks in CO ₂ from gas/liquid separator (8.12). The two phase liquid comprising lubricant and a dilute solution of CO ₂ in propylene carbonate is passed to a centrifugal separator (8.15) where the less dense lubricant is separated from the more dense solution. The lubricant is returned the compressor sump for lubricating the bearings and rotor. A small quantity of the lubricant becomes entrained in the compressed gas flow and is thus driven from the compressor. The solution is pumped from the centrifugal separator (8.15) by the liquid pump (8.13) into jet mixer (8.14) where it meets the compressed gas from compressor (8.2). The jet mixer comprises channels in the form of letter "T" whose diameters are smaller than the diameters of the gas and liquid feed pipes and the exit pipe for the gas/liquid mixture. The pressure drop across the mixer is sufficient to create a large mixing power input per unit volume at the point where the arms of the "T" meet. In this embodiment the liquid is fed into stem of the "T" and the gas into one side of the crossbar.

The compressed gas/solution is discharged into the absorber (8.7) where it is progressively cooled by transferring heat to the air stream (8.8) forced over absorber (8.7) by fan

(8.9). The air flow is essentially counter-current to the working fluid flow. As the CO ₂ and solution travels along the absorber (8.7) progressively more gas dissolves in the liquid until it reaches the expansion valve (8.6) where the gas is mainly or wholly in solution. The solution plus any residual gas travels through the expansion valve where adiabatic flashing of the CO ₂ from the solution commences and cools the solution/gas mixture. The mixture enters the adsorber (8.3) where it is progressively heated by the airflow (8.4) forced over adsorber (8.3) by a fan (8.5), while the airflow becomes progressively cooled. The air flow (8.4) is essentially counter-current to the working fluid flow. The increasing temperature of the solution in the desorber causes more CO ₂ to desorb. After exiting the desorber, the gas/oil mixture re-enters the gas/oil separator (8.12) to complete the circuit. The system shown in Figure 8 maintains the room at 20 ⁰ C with an external temperature up to 35 ⁰ C.

Figure 11 shows a ventilation system designed to ventilate and air condition room (1 1.1) comprising a heat pump to recuperate coolth from outgoing stale air and transfer it to

incoming fresh air. The purpose of the device is to provide a combination of ventilation and air conditioning to maintain the temperature of a room at 10 to 25 K below temperature external to the room by recuperating coolth from the outgoing air and transferring it to incoming fresh air. A further advantage of this device is that it allows the stale air exhaust and fresh intake to be positioned sufficiently far from each other to prevent stale air being sucked back into the building. The device comprises a scroll compressor (11.2) capable of simultaneously compressing a blend of HFC 134a (40 wt%), HFC 125 (48 wt%) and HFC 245fa (12 wt%) where the percentages refer to the masses of the components present. ISO 32 grade polyol ester (POE) is used as a lubricant. A condenser (11.3) is also provided. An expansion valve (1 1.6) drops the refrigerant blend from a higher pressure in the condenser to a lower pressure in an evaporator (11.7). A blower (11.5) sucks stale air from the room and blows the air over the condenser (1.3). A blower (1 1.9) sucks outside fresh air over the evaporator (11.7) and blows it into the room. Control instrumentation monitors the operation of device and controls the expansion valve.

Figure 11 shows a cross-section through a room includes a door (11.10) a window

(11.11). Scroll compressor (11.2) sucks in refrigerant blend vapour and compresses it from a lower pressure of the evaporator to a higher pressure of the condenser. After discharge from the compressor the hot higher pressure, super-heated gas enters the condenser. In the first part of the condenser the gas is desuperheated by the transfer of heat to the outgoing stale air until it reaches its dew point. In the second part of the condenser the HFC blend is transformed from gas to liquid with the removal of heat until it reaches its bubble point. In the third part of the condenser the liquid is sub-cooled by the transfer of heat to stale air leaving the room. The stale air flow driven by blower (11.5) is essentially counter-current to the refrigerant flow in the condenser. The liquid refrigerant blend then passes through the expansion valve where part is evaporated adiabatically reducing its temperature so that it enters the evaporator as a mixture of vapour and liquid at a lower pressure and temperature. As the refrigerant blend flows along the evaporator it progressively evaporates at increasing temperature until it reaches its bubble point with the absorption of heat from the incoming fresh air. In the second part of the evaporator further heat transfer from the fresh air superheats the refrigerant vapour by 5 K. The fresh air flow driven by blower (11.9) is essentially counter-current to the refrigerant flow in the evaporator.

After exiting the evaporator the refrigerant vapour mixture enters the suction port of the compressor to complete the circuit. The system shown in Figure 11 is intended to both ventilate the room with at least four air changes per hour and to maintain the room at 20 ⁰ C with an external temperature of up to 40 ⁰ C.

In another embodiment of this invention a heat pump is employed to heat a room and maintain its temperature at 10 to 30 K above the external temperature. The heat pump is represented iri Figure 12 which shows a cross section through a room which can be entered by a door (12.10) and is lit by a window (12.11). The heat pump comprises a scroll compressor (12.2) capable of simultaneously compressing a blend of HFC 134a (40 wt%), HFC125 (48 wt%) and HFC 245fa (12 wt%) where the percentages refer to the masses of the components present. ISO 32 grade polyol ester (POE) lubricant is added to the blend. A condenser (12.3) is provided. An expansion valve (12.6) reduces the refrigerant blend from a higher pressure in the condenser to a lower pressure in an evaporator (12.7). A blower (12.5) draws fresh air from the outside and blows over condenser (12.3); a blower (12.9) which sucks stale air from the room over 12.3 and blows it outside. Control instrumentation which monitors the operation of device and adjusts the expansion valve.

The scroll compressor (12.2) sucks in refrigerant blend vapour and compresses it from a lower pressure of the evaporator to a higher pressure of the condenser. After discharge from the compressor the hot higher pressure, super-heated gas enters the condenser. In the first part of the condenser the gas is desuperheated by the transfer of heat to the incoming fresh air until it reaches its dew point. In the second part of the condenser the HFC blend is transformed from gas to liquid with the removal of heat until it reaches the bubble point. In the third part of the condenser the liquid is sub-cooled by the transfer of heat to fresh air being sucked from outside. The fresh air flow driven by blower (12.5) is essentially counter-current to the refrigerant flow in the condenser. The liquid refrigerant blend then passes through the expansion valve where part is evaporated adiabatically reducing its temperature so that it enters the evaporator as a mixture of vapour and liquid at a lower pressure and temperature. As the refrigerant blend flows along the evaporator it progressively evaporates at increasing temperature with the absorption of heat from the outgoing stale air until it reaches its bubble point. In the second part of the evaporator further heat transfer from the stale air superheats the refrigerant vapour by 5 K by. The stale air flow driven by blower (12.9) is essentially counter-current to the refrigerant flow in the evaporator.

The system shown in Figure 12 is intended both to ventilate the room with at least 4 air changes per hour and to maintain the room at 20 ⁰ C with an external temperature down to -20 ⁰ C.

In an alternative embodiment of this invention a heat pump is configured such that air flows across the evaporator and the condenser can be reversed. In cooling mode the air flows

over the evaporator and condenser are in the directions shown in Figure 1 1. In heating mode the air flows are in the directions shown in Figure 12.

In a further embodiment of this invention an air-to-air recuperative coolth exchanger is combined with a heat pump containing a wide glide HFC blend which work together to recover heat from outgoing stale air and transfer it to incoming fresh air. Figure 13 shows this device installed to cool a room (13.1), maintaining at a temperature of 20 C with an external temperature of 35 ⁰ C. The device comprises a compressor (13.2); a condenser (13.3); an expansion valve (13.6); a blower (13.5); an evaporator (13.7); a suction line/liquid line heat exchanger (13.10) and a blower (13.9). A refrigerant blend comprising HFC 134a / HFC 125 / HFC 32 in the ratio 0.21/0.21/0.58 and an ISO 32 grade polyol ester compressor lubricant. An air- to-air recuperative coolth exchanger (13.11) and a control system which operates the expansion valve so that vapour leaving the evaporator has a selected superheat, are also provided.

Lower pressure refrigerant enters the compressor suction port and after compression is discharged as superheated gas at a higher pressure. In the first section of the condenser the high pressure gas is de-superheated until it reaches it dew point with the heat being transferred to the counter-current stale air stream (13.4) just before rejection to external environment. In the second section of the condenser the refrigerant is progressively condensed until it reaches its bubble point with its latent heat transferred to the stale air stream. Because the ratio of the components comprising the refrigerant blend are specially selected to condense over a temperature range the condensation temperature progressively falls. In the third section of the condenser the liquid refrigerant is subcooled again by transfer of heat to the stale air flow. The liquid refrigerant then passes through the suction line/liquid line heat exchanger where heat from higher temperature, higher pressure, liquid refrigerant is transferred to lower pressure, lower temperature refrigerant vapour. The liquid refrigerant then enters expansion valve where its pressure drops from a higher value in the condenser to a lower value in the evaporator. The pressure drop results in a rapid, adiabatic evaporation of a portion of the liquid refrigerant generating a cold mixture of vapour and liquid refrigerant. This process is commonly called "flashing". In the first section of the evaporator the cold lower pressure refrigerant is progressively heated absorbing heat from the incoming, counter-current, fresh air flow (3.8) which is progressively cooled. Because the ratio of the components comprising the refrigerant blend are specially selected to evaporate over a temperature range the evaporation temperature progressively rises as the refrigerant passes along the evaporator. In the second section of the

evaporator the refrigerant vapour is superheated by transfer of heat from the fresh air stream. After exiting the evaporator the refrigerant vapour enters the suction line /liquid line heat exchanger, where it is heated by the counter-current flow of liquid refrigerant from the condenser, and re-enters the suction port of the compressor to complete the circuit. In the system shown in Figure 13 the condenser is located in stale air stream expelled from the air-to-air, coolth recuperator. The evaporator is located in fresh air stream expelled from the air-to-air, coolth recuperator.

In another embodiment of this invention a heat pump, shown in Figure 14, is used simultaneously to generate a stream of cold water for air conditioning a room (14.1), which can be entered by a door (14.10) and is lit by a window (14.11). A hot water supply is provided for cleaning purposes. The system comprises a compressor (14.2); an evaporator (14.3); an expansion valve (14.4); a condenser (14.5) and a water circulation pump 14.6). A heat exchanger (14.9) located in the room (14.1). A hot water storage tank (14.12); a water/glycol circulation pump (14.13); a water-to-water heat exchanger (14.14); a suction line / liquid line heat exchanger (14.15) are provided. The refrigerant comprises R125/R245fa in the ratio ; plus an ISO grade 32 compressor lubricant. A control system enables the user to set the required operation of the heat pump. The control system comprises an inverter to vary the rotational speed of the compressor and a feed back mechanism that can change the aperture of the expansion valve to vary the quantity of refrigerant passing from the condenser to the evaporator.

Lower pressure refrigerant vapour enters the suction port of the compressor and is compressed to a higher pressure. Higher temperature, higher pressure superheated refrigerant gas is discharged from the compressor into the condenser. In the first section of the condenser the high pressure gas is de-superheated until it reaches it dew point with the heat being transferred to a counter-current water flow that is pumped around a circuit comprising the water storage tank and the condenser by pump 14.6. In the second section of the condenser the refrigerant is progressively condensed until it reaches its bubble point with its latent heat transferred to the counter-current water flow. Because the ratio of the components comprising the refrigerant blend are specially selected to condense over a temperature range the condensation temperature progressively falls. In the third section of the condenser the liquid refrigerant is subcooled again by transfer of heat to the counter-current water flow. The liquid refrigerant then passes through the suction line/liquid line heat exchanger where heat from higher temperature, higher pressure, liquid refrigerant is transferred to lower pressure, lower temperature refrigerant vapour. The liquid refrigerant then enters expansion valve where its pressure drops from a higher value in the

condenser to a lower value in the evaporator. The pressure drop results in a rapid, adiabatic evaporation of a portion of the liquid refrigerant generating a cold mixture of vapour and liquid refrigerant. In the first section of the evaporator the cold lower pressure refrigerant is progressively heated absorbing heat from a counter-current water/glycol flow which is progressively cooled. The water/glycol flow is pumped by pump (4.13) around a circuit comprising the condenser and the heat exchanger (14.9), where the water/glycol cooled by the evaporator by absorbing heat from air in the room. Because the ratio of the components comprising the refrigerant blend are specially selected to evaporate over a temperature range the evaporation temperature progressively rises as the refrigerant passes along the evaporator. In the second section of the evaporator the refrigerant vapour is superheated by transfer of heat from the counter-current water/glycol. After exiting the evaporator the refrigerant vapour enters the suction line /liquid line heat exchanger, where it is heated by the counter-current flow of liquid refrigerant from the condenser, and re-enters the suction port of the compressor to complete the circuit. Cold water from the mains water supply (14.8) passes through the water-water heat exchanger (14.14) where it is heated and exits as a hot water supply (14.7) for cleaning purposes. The hot water flow from the condenser enters the water storage tank at the top and the cold water feed to the condenser is drawn from the bottom. The mains water enters at the bottom of the exchanger (14.14) and leaves at the top. This arrangement provides a supply of hot water at essentially constant temperature although the water storage tank may contain denser, cold water in its lower part with lighter hot water in its upper part.

In an alternative embodiment of this invention the intermediary heat exchanger (14.14) is omitted.

Heat, known as geothermal energy, generated deep in the Earth by nuclear fission, flows outwards to the planet's surface where it is lost by radiation into space. In a further embodiment of this invention a heat pump containing a wide glide refrigerant blend may be employed to intercept geothermal energy and pump it to a higher temperature. Figure 5 represents a heat pump which uses geothermal energy as a heat source to heat a room (15.1) which can be entered by a door (15.10) and is lit by a window (15.11). A dotted line (15.8) indicates the surface of the ground.

The system comprises a compressor (15.2); an evaporator (15.3); an expansion valve

(15.4); a condenser (15.5) and a water circulation pump (15.6). A heat exchanger (15.9) is located in the floor of the room. A heat exchanger coil (15.12) containing a water/glycol mixture 100m in length is located 1.5 m below the ground surface. A water/glycol circulation pump

(15.13); a suction line / liquid line heat exchanger (15.15) are provided. The refrigerant comprises R744 (0.7 wt%), R125 (20 wt%), R32 (20 wt%) and R134a (53 wt%) plus an ISO grade 32 compressor lubricant. A control system comprises a control panel, which allows a user to select a room temperature, an inverter to vary the rotational speed of the compressor and a feed back mechanism that can change the aperture of the expansion valve to vary the quantity of refrigerant passing from the condenser to the evaporator.

Lower pressure refrigerant vapour enters the suction port of the compressor and is compressed to a higher pressure. Higher temperature, higher pressure superheated refrigerant gas is discharged from the compressor into the condenser. In the first section of the condenser the high pressure gas is de-superheated until it reaches it dew point with the heat being transferred to a counter-current water flow that is pumped around the under-floor water circuit and the condenser by pump (15.6). In the second section of the condenser the refrigerant is progressively condensed until it reaches its bubble point with latent heat also transferred to the counter-current water flow. Because the ratio of the components comprising the refrigerant blend are specially selected to condense over a temperature range the condensation temperature progressively falls. In the third section of the condenser the liquid refrigerant is subcooled, again by transfer of heat to the counter-current water flow. The liquid refrigerant then passes through the suction line/liquid line heat exchanger where heat from higher temperature, higher pressure, liquid refrigerant is transferred to lower pressure, lower temperature refrigerant vapour. The liquid refrigerant then enters expansion valve where its pressure drops from a higher value in the condenser to a lower value in the evaporator. The pressure drop results in a rapid, adiabatic evaporation of a portion of the liquid refrigerant generating a cold mixture of vapour and liquid refrigerant. In the first section of the evaporator the cold lower pressure refrigerant is progressively heated by absorbing heat from a counter-current water/glycol flow, which is progressively cooled. The water/glycol flow is pumped by pump (15.13) around a circuit comprising the condenser and the heat exchanger coil (15.12), where the water/glycol cooled by the evaporator absorbs geothermal heat (15.7), its temperature progressively rising as it flows through the coil. Because the ratio of the components comprising the refrigerant blend are specially selected to evaporate over a temperature range the evaporation temperature progressively rises as the refrigerant passes along the evaporator. In the second section of the evaporator the refrigerant vapour is superheated by transfer of heat from the counter-current water/glycol. After exiting the evaporator the refrigerant vapour enters the suction line/liquid line heat exchanger, where it is heated by the counter-current flow of liquid refrigerant from the condenser, and re-enters the suction port of the compressor to complete the circuit.

A heat pump may be used to produce a supply of hot water from a lower temperature heat source for washing and cleaning applications. Convenient heat sources include but are not limited to flows of air and mains water, solar energy and geothermal energy.

Visible and infra-red radiation from the Sun is absorbed at the Earth's surface and largely re-emitted as longer wave-length heat radiation. Although amount of incident radiation varies significantly according to time of year and latitude, on average over a year the amount of radiant energy radiation received and re-emitted is essentially the same. A solar, thermal energy collector may intercept incoming radiation, before it is absorbed by the Earth's surface and may use it to heat a heat transfer fluid, for example a water/glycol mixture, which in turn may be used to heat water for washing and cleaning applications. When radiation levels are low, for example during cloudy weather or during winter months, the temperature of water may not be high enough for cleaning and washing applications. The water temperature may be boosted by supplementary heat source, for example a mains-powered electrical heater or by burning fossil fuel, for example gas. Preferably the temperature boost is provided by a heat pump, which consumes less energy and thus creates less carbon dioxide to achieve the required temperature than a mains electrical heater or a direct fossil heat source.

Figure 6 shows a hot water heat pump using a solar thermal energy panel as a temperature heat source and heated by solar radiation (16.1). The system comprises a compressor (16.2); an evaporator (16.3); an expansion valve (16.4); a condenser (16.5); a water circulation pump (16.6); a hot water supply pipe (16.7); a mains water supply pipe (16.8); a solar, thermal energy collector (16.9); a water/glycol circulation pump (16.10); a hot water storage tank (16.12) and a suction line / liquid line heat exchanger (16.11). The refrigerant comprises R125 (55 wt%) and R236ea (45 wt%) plus an ISO grade 32 compressor lubricant. A control system comprises a control panel which allows a user to select a room temperature and a feed back mechanism that can change the aperture of the expansion valve to vary the quantity of refrigerant passing from the condenser to the evaporator.

Lower pressure refrigerant vapour enters the suction port of the compressor and is compressed to a higher pressure. Higher temperature, higher pressure superheated refrigerant gas is discharged from the compressor into the condenser. In the first section of the condenser the high pressure gas is de-superheated until it reaches its dew point with the heat being transferred to a counter-current water flow that is pumped from the bottom of the hot water storage tank by pump (6.6). In the second section of the condenser the refrigerant is progressively condensed

until it reaches its bubble point with latent heat also transferred to the counter-current water flow. Because the ratio of the components comprising the refrigerant blend has been specially selected to condense over a temperature range the condensation temperature progressively falls. In the third section of the condenser the liquid refrigerant is subcooled, again by transfer of heat to the counter-current water flow. The hot water flow leaving the condenser returns to the top of the hot water storage tank. The liquid refrigerant passes through the suction line/liquid line heat exchanger where heat from higher temperature, higher pressure, liquid refrigerant is transferred to lower pressure, lower temperature refrigerant vapour. The liquid refrigerant then enters expansion valve where its pressure drops from a higher value in the condenser to a lower value in the evaporator. The pressure drop results in a rapid, adiabatic evaporation of a portion of the liquid refrigerant generating a cold mixture of vapour and liquid refrigerant. In the first section of the evaporator the cold lower pressure refrigerant is progressively heated by absorbing heat from a counter-current water/glycol flow, which is progressively cooled. Because the ratio of the components comprising the refrigerant blend are specially selected to evaporate over a temperature range the evaporation temperature progressively rises as the refrigerant passes along the evaporator. In the second section of the evaporator the refrigerant vapour is superheated by transfer of heat from the counter-current water/glycol. After exiting the evaporator the refrigerant vapour enters the suction line/liquid line heat exchanger, where it is heated by the counter- current flow of liquid refrigerant from the condenser, and re-enters the suction port of the compressor to complete the circuit. The cold water/glycol mixture leaving the evaporator is pumped back to the thermal solar energy collector where it is reheated by the solar radiation.

A temperature gradient exists across the walls of a cooled enclosure, for example a commercial freezer, refrigerated container, domestic refrigerator or an air conditioned room. A heat flow down the temperature gradient tends to cause temperature of the cooled enclosure to rise. A refrigeration or air conditioning unit removes heat from a cooled enclosure via a cold heat exchanger thus maintaining a lower temperature inside the enclosure relative the temperature of it surroundings. The cold heat exchanger may be an evaporator in which a refrigerant changes state from liquid to vapour absorbing its latent heat of vaporisation from the enclosure. Alternatively the heat exchanger may carry a flow of a heat transfer fluid, which may be called a secondary refrigerant, for example water, a brine or a water/glycol mixture. The cold, heat transfer fluid is heated as it absorbs heat from the cooled enclosure and is pumped to the evaporator of a heat pump where it re-cooled before recycling to the cold heat exchanger in the cooled enclosure. Whether an enclosure is cooled directly by a heat pump evaporator or

indirectly by a secondary refrigerant circuit the process is thermodynamically inefficient because all the heat being removed is pumped from a lower temperature in the enclosure to a higher external temperature. In an alternative embodiment of this invention greater energy efficiency may be achieved by embedding heat exchangers in at least one wall of a cooled enclosure.

Figure 17 shows one design for an embedded exchanger/heat pump system for maintaining a cold room (17.1) for storing perishable foodstuffs at 4 ⁰ C. The system comprises a compressor (17.2); an evaporator (17.3); an expansion valve (17.4); a condenser (17.5); a water circulation pump (17.6); heat exchangers (17.9a-d) embedded in the walls of the cold room carrying a water/glycol secondary refrigerant from and to the evaporator (17.3); a water/glycol circulation pump (17.10); a hot water storage tank (17.12); a suction line / liquid line heat exchanger (16.11). The refrigerant comprises R134a (42 wt%), R125 (25 wt%), R32 (25 wt%), and CO ₂ (8 wt%) plus an ISO grade 32 compressor lubricant. A control system comprises a control panel which allows a user to select a room temperature and a feed back mechanism that can change the aperture of the expansion valve to vary the quantity of refrigerant passing from the condenser to the evaporator. For clarity only pipe connections between heat exchanger (17.9a) and the heat pump evaporator (17.3) are shown. The pipe connections between the evaporator and (17.9b-d) are in parallel with the evaporator connections to (17.9a), and with each other. Heat exchangers (17.9a-d) each comprise three sections. For (17.9a) these are designated (17.9aA), (17.9aB) and (17.9aC). Section (17.9aA) essentially is part of the inner wall of the cooled room. Sections (17.9aB) and (17.9aC) are located within the insulation that comprises the wall of the room.

Lower pressure refrigerant vapour enters the suction port of the compressor and is compressed to a higher pressure. Higher temperature, higher pressure superheated refrigerant gas is discharged from the compressor into the condenser. In the first section of the condenser the high pressure gas is de-superheated until it reaches its dew point with the heat being transferred to a counter-current water flow that is pumped from the bottom of the hot water storage tank by pump (17.6)

In the second section of the condenser the refrigerant is progressively condensed until it reaches its bubble point with latent heat also transferred to the counter-current water flow. Because the ratio of the components comprising the refrigerant blend are specially selected to condense over a temperature range the condensation temperature progressively falls. In the third section of the condenser the liquid refrigerant is subcooled, again by transfer of heat to the counter-current water flow. The hot water flow leaving the condenser returns to the top of the hot

water storage tank. The liquid refrigerant passes through the suction line/liquid line heat exchanger where heat from higher temperature, higher pressure, liquid refrigerant is transferred to lower pressure, lower temperature refrigerant vapour. The liquid refrigerant then enters expansion valve where its pressure drops from a higher value in the condenser to a lower value in the evaporator. The pressure drop results in a rapid, adiabatic evaporation of a portion of the liquid refrigerant generating a cold mixture of vapour and liquid refrigerant. In the first section of the evaporator the cold lower pressure refrigerant is progressively heated by absorbing heat from a counter-current water/glycol flow, which is progressively cooled. Because the ratio of the components comprising the refrigerant blend are specially selected to evaporate over a temperature range the evaporation temperature progressively rises as the refrigerant passes along the evaporator. In the second section of the evaporator the refrigerant vapour is superheated by transfer of heat from the counter-current water/glycol. After exiting the evaporator the refrigerant vapour enters the suction line/liquid line heat exchanger, where it is heated by the counter- current flow of liquid refrigerant from the condenser, and re-enters the suction port of the compressor to complete the circuit. The cold water/glycol mixture leaving the evaporator is passes to the section of the insulation embedded heat exchanger (17.9a) in contact with the inner wall of the refrigerated room where absorbs heat from warmer articles and heat being conducted, convected and radiated from the room surroundings which may be at a higher temperature than the interior. The circulating water/glycol then successively flows into the middle section of exchanger (17.9a) where it may absorb more heat flowing inwards from the surroundings thus further increasing its temperature before entering into the outer section of exchanger (17.9a) where it may absorb more inward flowing heat. The water/glycol is driven by pump (17.10) back to the condenser to complete the water/glycol circuit.

This embodiment is especially energy efficient because it recovers coolth as it flows out through the insulation. Inward flowing heat may be captured at progressively higher temperatures as the water glycol mixture flows outwards by passing successively through the three sections that comprise heat exchanger (17.9a). The water/glycol mixture enters (17.9aA) at -3 ⁰ C and leaves at +1 ⁰ C in order to maintain the room at the desired temperature of 4 ⁰ C; it then enters (17.9aB) at 1 ⁰ C and leaves at + 7 ⁰ C; and finally it enters (17.9C) at 7 ⁰ C and leaves at 12 ⁰ C. Heat is pumped from (17.9aA) at a temperature lower than the room temperature, but is pumped at temperature higher than the room temperature from (17.9aB) and from (17.9aC). The energy input to the heat pump compressor required to pump the heat from (117.9aC) is less than the energy input required to pump heat from (17.9aB), which in turn is less than the energy input required to pump heat from (17.9aA). In an existing cooled the heat exchanger carrying the

circulating water/glycol mixture may be located inside the room and will absorb all the heat the passing through the external faces of the walls at a temperature lower than -3 ⁰ C. This is necessarily requires more energy than the embodiment of the present invention represented in Figure 17.

A further advantage of this embodiment is the utilisation of the heat rejected via the condenser to generate water for cleaning and washing. This may avoid, the emission of more CO ₂ from the burning of fossil fuel, e.g. gas, in a separate system to heat the water. Compared to a single refrigerant, e.g. R22, or an azeotrope, e.g. R507, or a near-azeotrope, e.g. R410A, the large temperature glide in the condenser that may be obtained by using a zeoptrope enables the system to produce water at the desired high temperature while allowing a lower operating pressure. If the heat from the condenser is at too low a temperature to be used for cleaning and washing its temperature may require boosting by burning fossil fuel thus contributing more to reduction of global warming than the embodiment of the present invention represented in Figure 17.

Figure 8 shows the design for an embedded heat exchanger section such as (7.9aB). The section comprises a set of conduits (18.1) through which a heat transfer liquid (secondary refrigerant, e.g. a water/glycol mixture) is pumped; a metal (e.g. aluminium) sheet (18.2) which is parallel to the inner and outer faces of the cooled enclosure and in thermal contact with the conduits; a water/glycol entry manifold (18.3) which distributes lower temperature secondary refrigerant from the heat pump evaporator to conduits (18.1); and a water/glycol exit manifold (18.4) which collects higher temperature secondary refrigerant from conduits (18.1) for return to the heat pump evaporator.

When the heat exchanger section is installed in a wall of a cooled enclosure, a sheet (18.2) may capture heat transmitted from the outer face of the wall and conduct it to the conduits (18.1) and hence to secondary refrigerant which transports the captured heat to the heat pump evaporator.

In a preferred aspect of the embodiment of this invention represented by Figure 17, the wall embedded heat exchanger, shown in Figure 9a, comprises a metal sheet which is bent to provide contiguous sections (19.2a, 19.2b, and 19.2c); a set of conduits (19.1) in thermal contact with the sheet through which passes a secondary refrigerant; a water/glycol entry manifold (19.3) which distributes lower temperature secondary refrigerant from the heat pump evaporator to conduits (19.1); and a water/glycol exit manifold (19.4) which collects higher temperature

secondary refrigerant from conduits (19.1) for return to the heat pump evaporator. Figure 19b represents a cross-section through a heat exchanger, such as that shown in Figure 19a, embedded in a wall of a cooled enclosure. Section (19.1) of the embedded heat exchanger essentially comprises the inner face of the wall of cooled enclosure. Sections (19.2) and (19.3) are inclined to the inner and outer faces of the wall such that the temperature of the secondary refrigerant at any point may differ from the wall at that point by not more than 10 K, preferably not more than 5 K, and more preferably by not more than 3 K.

In a preferred embodiment of this invention the glycol secondary refrigerant is omitted.

The wide glide blend refrigerant passes through the embedded heat exchanger. Without being limitative a systems of this design is represented in Figure 20. The same component numbering of Figure 17 is retained in Figure 20 for clarity. The expansion valve (20.4) is sited near the entry to innermost section of embedded heat exchanger.

Examples

Examples 1 to 4 An air conditioning unit as shown in Figure 1 is operated as described previously to cool room 1.1 with the input parameters given in upper part of Table 2. The COP and other performance results are summarised in lower part of the Table 2.

Examples 5 to 8

A heat pumping unit as shown in Figure 2 is operated to heat room 2.1 as described previously with the input parameters given in upper part of Table 2. The COP and other performance results are summarised in lower part of the Table 3.

Examples 9 to 12

A heat pumping unit as shown in Figure 3 was operated to heat a room as described previously by extracting heat from the ground with the input parameters given in upper part of Table 4. The COP and other performance results are summarised in lower part of Table 4.

Examples 13 and 14

The performance of an air-to-air heat pump, which includes a recuperative heat exchanger as represented in Figure 13 and an hermetic compressor, was evaluated using the

NIST CYCLE D program with a refrigerant blend comprising R32 (21 weight %), R125 (21 weight %) and 134a (58 weight %). The modelling conditions are given below in Table 5a. For comparison the performance of R22 was also evaluated with the operating conditions selected to

give the essentially the same evaporator inlet and outlet temperatures as for the blend. R22 is an ozone depleting refrigerant, widely used in heat pumps, which is being phased-out in accordance with the Montreal Protocol.

The results of analysing the performance of the blend in the air-conditioning unit are shown in Table 5b, together with the results for R22. It is clear that the blend in Example 13 is superior to R22 in having a higher cooling and heating COP and thus being more energy efficient, and having a higher cooling and heating capacity. The blend operates at similar compressor discharge pressure to R22 and has a lower compressor discharge temperature, which will make the air conditioning unit more reliable.

Examples 14 to 16

The performance of a ground heat pump, as represented in Figure 13 and an hermetic compressor, was evaluated using the NIST CYCLE D program, with two refrigerant blends. The blend in Example 16 comprises CO ₂ (7 weight %), R32 (20 weight %), R125 (20 weight %) and 134a (55 weight %). The blend in Example 17 comprises ethane (25 weight %). propane (65 weight %) and isobutane (10 weight %). The modelling conditions are given in Table 6a. For comparison the performance of R410A, a 50/50 weight blend of R32 and Rl 25, was also evaluated with the operating conditions selected to give essentially the same evaporator and condenser inlet and outlet temperatures as for the blend (Example 15). R410 is a commercially available blend having essentially zero glide, which is replacing R22 in heat pump applications.

The results of analysing the performance of the blends in the ground heat pump are shown in Table 6b, together with the results for R410A. It is clear that both the blends in Examples 16 and 17 are superior to R410 in having a higher cooling and heating COPs and thus being more energy efficient, and having a higher cooling and heating capacities. The blends operate at lower compressor discharge pressures than R22 and have lower compressor discharge temperatures, which will make the heat pump more reliable.

Examples 18 to 21

The performance of a refrigerator, as represented in Figure 20 equipped with an hermetic compressor, was evaluated using the NIST CYCLE D program, with three refrigerant blends. The refrigerator was designed both to store frozen food and to supply hot water for washing and cleaning. The blend in Example 19 comprises CO ₂ (7 weight %), R32 (20 weight %), Rl 25 (20 weight %) and 134a (55 weight %). The blend in Example 20 comprises ethane (25 weight %). propane (65 weight %) and isobutane (10 weight %). The blend in Example 21

comprises ethane (25 weight %). propane (65 weight %) and isobutane (10 weight %). The modelling conditions are given in Table 7a. For comparison the performance of R404A, a blend of R143a ( 52 wt %) R125 (42 wt%) and R134a (4 wt %), was also evaluated with the operating conditions selected to give essentially the same evaporator and condenser inlet and outlet temperatures as for the blend (Example 15). R410 is a commercially available blend having essentially zero glide, which is replacing R22 in heat pump applications.

The results of analysing the performance of the blends in the ground heat pump are shown in Table 7b, together with the results for R410A. It is clear that both the blends in Examples 19 to 20 are superior to R410 in having a higher cooling and heating COPs and thus being more energy efficient. Blends 19 and 21 also have higher cooling capacities than R410A. Blends 19 to 21 all have lower discharge temperatures and pressures than R410A , which will make the refrigeration unit more reliable. Blends 19 and 20 are especially advantageous in this respect.

Example 22

A heat pump unit, represented by Figure 10, was constructed to provide a stream of cold water from a mains water supply. The working fluid charged to the system comprised Rl 25 (48 wt%), Rl 34a (40 wt%) and R245fa (12 wt%). The lubricant was Emkarate™ RH32L (0.650 kg). The refrigerant circuit comprised: an Hitachi hermetic scroll compressor, model ZS7516S1 nominal refrigerating capacity 1 horse power (10.1); a Packless coaxial condenser, model COAX-2150-H, nominal capacity 1.5 refrigeration ton (10.2); expansion device consisting of two 0.25 inch adjustable needle valves in series (10.4); a Packless coaxial evaporator, model CHAX-3200-H, nominal capacity 2.0 refrigeration ton (10.3); Titan Industries, "beige", turbine water flow meter (10.5) for measuring the water flow to the condenser; Titan Industries, "beige", turbine water flow meter (10.6) measuring the flow of water to the evaporator; a manual control valve (10.8) for regulating the total flow of mains water (10.7) to the heat exchangers; a manual valve (10.9) for regulating the flow of water to the condenser; a manual flow valve (10.10) to regulate the flow of water to the evaporator; an accumulator (10.11) to catch liquid refrigerant and lubricant exiting the evaporator and then feed them back to the compressor; a sight glass (10.12) to observe whether liquid refrigerant and lubricant was returning to the compressor; a sight glass (10.13) to observe whether a mixture of liquid and vapour was entering the expansion device; a 24 bar gauge relief valve (10.14) to release refrigerant if the maximum pressure in the equipment exceeded its normal operating pressure of up to 20 bar gauge; two transducers (10.24 and 10.26) to measure the pressures at the entry and exit of the condenser respectively; two

transducers (10.27 and 10.25) to measure the pressures at the entry and exit of the evaporator respectively; a thermocouple (10.17) to measure the temperature of the compressor casing; two thermocouples (10.19 and 10.15) to measure the temperature of the water entering and leaving the condenser respectively; two thermocouples (10.23 and 10.22) to measure the temperature to the water entering and leaving the evaporator respectively; two thermocouples (10.21 and 10.18) to measure the temperature of the refrigerant entering and leaving the evaporator respectively; two thermocouples (10.16 and 10.20) to measure the temperature of the refrigerant entering and leaving the evaporator respectively; electronic units to collect data from the thermocouples, flow meters and pressure transducers and feed it to a computer for monitoring and storage; and a meter in the electrical supply to the compressor to measure its power consumption. The evaporator was insulated to prevent heat gain from the surrounding air and to prevent water condensation on the surface.

After charging with 0.1 kg of lubricant to augment the 0.550 kg supplied in the compressor the equipment was pumped with a vacuum pump to remove air and the liquid refrigerant charged. The water flows to the condenser and evaporator were then started and the compressor switched on. The refrigerant flows through the sight glasses were observed. The flow of water through the condenser was adjusted using the manual regulating valves (10.8) and (10.9) until only liquid was entering the expansion device. The needle valves comprising the expansion device were adjusted until no liquid refrigerant was observed in site glass (10.12). The equipment was then allowed to operate until it reached essentially steady state and the water flow and expansion device was adjusted until the desired output water temperature of around 5 ⁰ C was obtained.

The output data from the instrumentation was saved on the computer. A representative set of 12 data points taken 1 second intervals was averaged and used to calculate the performance of the device. The duty of the evaporator was determined by multiplying the water flow rate, converted to kg/s, by the heat capacity of water and by the difference between the inlet and outlet water temperatures. The duty of the condenser was similarly calculated. The cooling COP was calculated by dividing the evaporator duty by the power input to the compressor. Although the device was designed as a chiller the heating COP was calculated by dividing the duty condenser duty by the compressor power input.

The condensing temperatures of the two refrigerants were calculated from the condenser pressures and their thermodynamic properties calculated using the NIST REFPROP program. For the blend the glide was calculated from the difference between its dew and bubble point

temperatures at the pressure of the condenser. The difference between the bubble point and the refrigerant exit temperature from the condenser provided the subcool. The difference between the compressor discharge temperature and the dew point gave the de-superheat.

Using the measured pressure of the blend in the evaporator and REFPROP the dew point of the blend in the evaporator was calculated.

The performance of Rl 34a was measured for comparison with that of the blend. Table 8 summarises the results obtained. The COP values show that the blend is 17% more energy efficient than 134a. The cooling duties show that the cooling capacity of the blend is 26% higher than that of R134a. A further advantage is that blend allows the compressor to operate 17.8 K lower than is observed with 134a which will enhance the compressor reliability.

Example 23

An absorption heat pump, represented by Figure 9, was constructed to provide a stream of warm water from a mains water flow. The absorbent was 1.5 kg Emkarate™ RH 32L, which also acted as the lubricant for the compressor. The rig was charged with CO ₂ gas such that the working pressure in the discharge pressure under operating conditions was 19.6 bara.

The equipment comprised: an Hitachi scroll compressor model ZS408451 rated at 0.5 horse power (9.1) to compress the CO ₂ gas; a flat plate desuperheating heat exchanger (9.2); an Alfabiz, coaxial, absorber heat exchanger rated at 5 kW (9.3); a 0.5 inch BSP needle valve as the expansion device (9.4); an Alfabiz, coaxial, absorber heat exchanger rated at 5 kW (9.5); liquid- gas separator (9.6); T-jet mixer (9.7); inline mixer (9.8); sight glasses (9.9 to 9.13 and 9.38); two Fluid-o-Tech TMFRl liquid metering pumps (9.14 and 9.15) to pump the absorbent; two Titan Industries, "grey", turbine water flow meters (9.16 and 9.17); pressure transducers (9.18 to 9.23); thermocouples (9.24 to 9.37); electronic units (not shown) to collect data from the thermocouples, flow meters and pressure transducers and feed it to a computer for monitoring and storage; valves to isolate the pumps (9.39 and 9.40); valves to manually adjust the water flows through the absorber and the desorber (9.41 and 9.42) and safety release valve set to 27 bar gauge (9.43).

The dashed lines in Figure 9 represent the water flow pipe the solid lines connecting the components represent the CO ₂ /absorbent flow pipes.

After charging the unit the water flows were adjusted to the desired values by valves 9.41 and 9.42, and the compressor and one liquid pump switched on. The expansion valve was adjusted

until the ratio of the suction and discharge pressures was ~2 and the unit was allowed to achieve essentially steady state operation. The output data from the instrumentation was then captured and stored.

The duty in kW of each heat exchanger was calculated by multiplying the water flow through it in kg/s by the heat capacity of water and the temperature difference of the water entering and exiting the heat exchanger. The heat removed from the compressed gas by the superheater (9.6) essentially returned the temperature of the gas back to its suction value and was approximately equal to the work done in compressing the gas. The heating COP of the unit was obtained by dividing the sum of the duties of the desuperheater and the absorber by the duty of the desuperheater. The operating conditions and results are summarised in Table 9. Twenty data sets taken at 1 second intervals when the unit was operating under essentially steady state conditions.

The average value for each of the parameters reported is shown in the Table 9.

Tables

Table l(a)

The evaporator and condensation temperatures are mid-point values.

Table l(b)

Table 2

Table 3

Table 4

Table 5a

Table 5b

Table 6b

Table 7b

Table 8

Table 9