GB2508614A - Refrigerant compositions and heat pump - Google Patents

Abstract

A refrigerant comprises a plurality of molecules forming a zeotropic mixture having a high critical temperature and a wide glide comprising a wide temperature difference between the temperature at the bubble point and the temperature at the dew point of the mixture at a particular pressure. The temperature difference between the bubble point and the dew point is preferably at least 7°C and the critical temperature is preferably at least 135°C. The refrigerant may comprise a first compound selected from cyclobutane and 1,2-butadiene and a second compound selected from cyclobutene, butane, trans-2-butene, 1,3-butadiene and disilane. Heat pumps implemented using the refrigerants are also described as well as a method of determining a refrigerant composition in a heat pump.

Description

Heating System and Method Thc present invention relates to the field of refrigerants, particularly refrigerants for use in a heat pump. The invention also relates to heat pumps themselves.

A heat pump extracts energy from the ambient environment and use electricity to upgrade this energy to a useffil temperature. Heat pumps can be used to increase or decrease the temperature of a target environment. In particular a heat pump can be used to heat a space such as a house or to cool such a space as an air conditioning system.

Hcat pumps can provide a sourcc of heating or cooling that has a low cnvironmcntal impact and in some situations, can provide a cost-effective alternative, in particular where they are used in place of oil-fired heating systems.

Ground-source hcat pumps (GSHP), or geothcrmal hcat pumps, usc the ground as a hcat source or heat sink to respectively heat or cool a target environment. In particular, in a heating application, the evaporator of the heat pump is coup'ed to the ground to extract heat energy from the ground and deliver it via a heat exchanger to the target environment, such as a house.

Typical OSI-IPs can output fluid at a temperature of 50°C, with a return temperature of 42°C. Howcvcr, many cxisting domcstic heating systems USC radiators or othCr forms of Low Temperature Hot Water (LTHW) systems that are designed to operate at a temperature of around 70°C. Further, an output temperature of 50°C is not sufficient to prevent the growth of legionella, so such a OSI-IP cannot be used to provide domestic hot water for storage without a secondary form of heating to boost the temperature, such as an emersion heater.

Sincc the output temperature of a GSHP is too low for direct usc in a convcntional radiator heating system, GSHPs are usually used together with an undertloor heating system, which requires a lower operating temperature. However, it is costly and inconvenient to retrofit an underfloor heating system into a property and it is also wasteful if the property already has a radiator heating systcm.

Therefore, it would be advantageous to improve existing GSHPs to enable them to interface directly with radiator heating systems and to enable them to be used to heat domestic hot water.

According to one aspect, there is provided a refrigerant comprising a plurality of molecules forming a zeotropic mixture having a high critical temperature and a wide glide comprising a wide temperature difference between the temperature at the bubble point and the temperature at the dcw point of the mixture at a particular pressure.

It has been appreciated that a ncw rcfrigerant having these propcrtics cnables a GSHP to operate within parameters that means it can be used for domestic heating purposes. In particular, a GSHP using the refrigerant may produce an output water temperature sufficient to enable it to interface directly with a wet radiator system. Such a GSHP may also produce water at a temperature sufficiently high to enable it to be used safely as a source of hot water, for example in a domestic hot water system.

As described in more detail below, a combination of refrigerant components that produces a mixture having both a high critical temperature and a wide glide can be used to provide a GSHP having a higher output water temperature than in existing systems while still having an efficiency at least comparable to, if not better than other GSHPs.

Preferably. the temperature difference between the bubble point and the dew point of the mixture is at least 7°C, preferably at least 8°C. Zeotropic designs use temperature glide to advantage, by adopting counter-current evaporator and condenser heat exchangers which allow co-current flow of, and heat transfer between, the refrigerant liquid and vapour mixtures.

Preferably the temperature difference between the bubble point and the dew point of the mixture is substantially equal to (preferably within 1°C of) the temperature change of the service fluid that is being heated as it passes through the condenser. By matching the refrigerant side temperature glide with the service fluid temperature change, counter flow design lowers discharge pressures and raises suction pressures, so that the compressor does less work to deliver the same amount of heating (or cooling). Power savings of up to 25% may be achieved relative to comparable single component reftigerants.

In preferred embodiments, the critical temperature is higher than the dew poillt temperature by at least half the difference between the dew point and bubble point temperatures and preferably by at least the difference between the dew point and bubble point temperatures.

As set out in more detail below, the use of a refrigerant with a critical temperature significantly higher than the dew point of the mixture can enable efficient operation of a heat pump using such a refrigerant.

In highly preferred embodiments, the rcfrigcrant has a critical temperature grcatcr than 135°C, preferably greater than 150°C, flirther preferably greater than 165°C.

Preferably at least one molecule, preferably all molecules in the refrigerant, have a critical temperature greater than 135°C, preferably greater than 150°C, further preferably greater than 165°C.

In prcferrcd cmbodimcnts, the temperature at thc dew point of the mixture at atmospheric pressure is less than 15°C, preferably less than 12°C. This can enable the refrigerant to be used in a ground-coupled system, in which the temperature of the ground is typically 10-12°C.

Preferably, the temperature at the bubble point of the mixture at atmospheric pressure is greater than 0°C. It can be advantageous to maintain the operation of a heat pump abovc 0°C so that water-based components can be used in conjunction with the heat pump without the water freezing.

Preferably, the temperature at the dew point of the mixture at a pressure of 10 bar (1,000,000 Pa) is greater than 65°C, preferably greater than 70°C.

In a specific, highly preferred embodiment, the refrigerant is designed to condense between 80°C and 72°C at a pressure of 10 to 11 bar. Preferably, such a refrigerant can be used to produce hot water at 74°C flow and 66°C return, giving an ayerage of 70°C in compliance with EN442.

In preferred embodiments, the plurality of molecules comprise hydrocarbon molecules.

Thcsc moiccuics arc naturally-occurring and havc a low cnvironmcntal impact if rclcascd into the environment, so provide advantages over existing refrigerants such as fluorocarbons, which are less environmentally friendly. Further, the use of hydrocarbon refrigerants can reduce wear and tear on the compressor, in particular due to lower refrigerant density and viscosity, thus increasing its operating life. Hydrocarbons also provide improved heat transfer characteristics compared to other refrigerants.

Further, there is also no acid corrosion duc to moisturc ingress whcn using hydrocarbon rcfrigcrants, whcrcas whcn using fluorocarbons, highly corrosivc hydrofluoric acid is formed when moisture enters the system.

Preferably there is no chemical interaction between components of the zeotropic mixture.

Prcfcrably, thc molcculcs cach comprisc onc or zcro fluorinc atoms. Such molccuics havc a low environmental impact, since they do not cause as large global warming effects as those including fluorine, nor thc biopcrsistant acid rain problems causcd whcn molecules containing thrcc or more fluorine atoms break down to form trifluoracetic acid (TFA).

In a preferred embodiment, the refrigerant comprises at least one molecule selected from a first group consisting of: cyclobutane and I,2-butadicne. Each of thcsc molecules has a very high critical temperature, so increases the average critical temperature of the mixture.

Furthcr, thcsc molecules havc a high normal boiling point, so contributc to thc widc glide properties of the zeotropic mixture.

In a preferred embodiment, thc rcfrigcrant compriscs at least onc molecule selected from a second group consisting of: cyclobutene, trans-2-butene, butane, 1,3-butadiene, disilane and fluoropropene. Each of these molecules has a low normal boiling point, together with a relatively high critical tcmpcraturc, so thcy can bc uscd to incrcasc thc glidc of thc mixturc while maintaining the critical temperature at a high level.

In a highly preferred embodiment, at least 80%, preferably at least 90% by weight of the rcfrigcrant compriscs a combination of molcculcs selccted from the first and sceond group.

Further components may also be added to further increase the temperature glide or to vary the pressure.

Preferably, the refrigerant comprises a combination of at least 25% by weight of a molecule S selected from the fir st group of molecules and at least 25% by weight of a molecule selected from the second group of molecules. Further preferably, the mixture comprises at least 30%, preferably at least 40% of a molecule selected from the first group and a molecule selected from the second group.

In onc embodiment, thc refrigerant comprises 40-60% by weight of I,2-butadiene and 60- 40% by weight of cyclobutene.

Preferably, the refrigerant comprises substantially equal proportions by weight of 1,2-butadiene and cyclobutene. Combining these molecules in substantially equal proportions (for example within 10% of each other within the total mixture) can provide a refrigerant having the desired properties of high critical temperature and wide glide.

In preferred embodiments, the refrigerant comprises a component having an odour. This can provide an additional safety feature to enable human-detection of any leaks and is particularly advantageous when the refrigerant is a flammable substance.

A further aspect provides a heat pump comprising: a ground-coupled evaporator; and a refrigerant according to the preceding aspect or any of its dependent features. Such a heat pump can provide output temperatures and efficiencies as described herein.

A skilled person would appreciate that the principles described herein could also be applied to an air-conditioning system in which heat can be extracted from a relatively cool target environment and expelled into a relatively hot heat sink. In particular, the refrigerants and heat pumps described herein provide particular advantage in very high ambient climates such as the Middle East where external design conditions in excess of 50°C are now being requested. Therefore, references herein to "heating" may refer to "cooling" in embodiments implementing air conditioning systems.

A further aspect provides a heat pump comprising: a ground coupled evaporator; a hydrocarbon refrigerant; the heat pump having a nominal average working temperature of at east 65°C. Preferably, the nominal average working temperature is 70°C. Broadly, the nominal average working temperature can be defined as the average of the output and input temperatures of the heat pump. For example, a heat pump with a nominal average working temperature of 70°C may output fluid at a flow temperature of 75°C with a fluid return temperature of 65°C. More specifically, the nominal average working temperature can be defined as the log mean average temperature between the output and the input of the heat pump.

In preferred embodiments, the heat pump comprises a condenser and the pressure of the refrigerant in the condenser of the heat pump is less than 20 bar (2,000,000 Pa), preferably less than 15 bar (1,500,000 Pa). Preferably, the pressure is around 10 bar (1,000,000).

Providing a refrigerant at relativ&y low pressures increases the critica' temperature for the refrigerant, enabling the heat pump to run more efficiently. Further, using lower pressures can reduce the likelihood of refrigerant leakage and reduce the need for the use of pressure vessels in the heat pump and during service operations.

Preferably, the heat pump further comprises means for mixing the refrigerant. The means is preferably provided within the closed ioop circulating the refrigerant and enables the zeotropic mixture to remain mixed even though refrigerant components evaporate and condense out of the mixture at different temperatures. The mixing means may comprise a passive mixing means such as a tube-in-tube mixing serpentine heat exchanger in which the outer tube pipe diameter is varied at every bend. The tubes carrying the refrigerant may themselves be designed to encourage turbulent flow within the refrigerant, hence encouraging mixing. This may be implemented for example using internally-ribbed or finned tubes or tubes with twisted raised surfaces on the intemal walls. Alternatively, or in addition, an active mixing means may be provided. For example, a turbine may be provided that could be driven by external electrical means or preferably is simply driven by the flow of the refrigerant fluid, the fluid being mixed as it passes through the blades of the turbine.

Preferably the heat pump further comprises pressure reduction means for reducing the temperature and pressure of the refrigerant and control means for controlling the operation of the pressure reduction means. Preferably, the pressure reduction means comprises an expansion device such as an expansion valve.

Preferably the heat pump further comprises a temperature sensor for determining the temperature of the refrigerant wherein the determined temperature is input into the control means.

Preferably the heat pump further comprises a compressor and the temperature sensor is arranged to measure the temperature of the refrigerant prior to the refrigerant entering the compressor.

The control means can be used to ensure the refrigerant is expanded to an appropriate temperature and pressure to optimise the running of the heat pump. For example, as the temperature of the target environment rises, less heat will be absorbed from the refrigerant by the service fluid so the refrigerant entering the pressure reduction means will be at a higher temperature and pressure. Therefore, the pressure reduction means will need to do more work on the refrigerant to lower it to an appropriate temperature and pressure for use in the evaporator.

Preferably, the heat pump further comprises a compressor and compressor control means for controlling the operation of the compressor.

Prcfcrably, the nominal output temperature of thc heat pump is grcatcr than 65°C, preferably greater than 70°C. Therefore, in operation, service fluid such as water can be output from the heat pump system above 65°C, preferably above 70°C, further preferably at the European standard EN442 temperatures of 75CC flow / 65CC return.

In some embodiments, the nominal input temperature of the heat pump is around 60°C, preferably substantially 65°C. This may be the temperature of the service fluid such as water returning to the heat pump system.

Preferably, the compressor control means controls the compressor based on the nominal input temperature and further preferably the nominal output temperature, of the heat pump.

S

For example, as the nominal input temperature approaches 75°C, the compressor control means may cause the compressor to slow or stop the pumping of refrigerant around the system, since the target environment has reached a temperature wherein it is not absorbing significant amounts of heat from the service fluid. Therefore, the target environment has reached the desired temperature. Further, the control means can slow or stop the compressor if the nominal output temperature of the heat pumps exceeds a specified value, such as 75°C. This can provide a safety cut-off to ensure that the service fluid does not rise above the specified value to prevent users being burnt by radiators or hot water heated by the heat pump.

The compressor control means preferably comprises an inverter to eontrol the capacity of the compressor.

The operation of the compressor and/or pressure reduction means may be controlled dynamically, in particular based on feedback obtained relating to the pressures and temperatures of operation of the system.

A further aspect provides a method of determining the composition of a refrigerant in a heat pump, the method comprising obtaining a measure of the thermal characteristics or the performance of the heat pump and analysing the measure to determine the relative proportions of components in the refrigerant of the heat pump.

This can enable refrigerant levels to be adjusted or topped up in a heat pump with a zeotropic refrigerant. Since a zeotropic refrigerant will distil into its constituent components as part of the evaporation/condensation cycle, a refrigerant leak may cause loss of one component of the mixture without loss of other components. The thermal characteristics or performance of the heat pump can be used to provide an indicator of which component of the refrigerant is too low or too high and the levels of components in the refrigerant mixture can be adjusted accordingly. For example, a drop in efficiency of the heat pump may indicate that the critical temperature of the refrigerant needs to be increased, so the refrigerant component with the higher critical temperature can be added. Similarly, a reduction in the temperature of the refrigerant exiting the evaporator may indicate that the refrigerant component with the lowest bubble point is too low and this may therefore be increased.

Preferably, the method includes providing an indication of a component of the refrigerant that is higher than a desired level. This can enable a service engineer to top up the refrigerant with the appropriate component to provide an appropriate zeotropic blend within

an acceptable range.

In a preferred embodiment, the heat pump may include a data collection system to collect and store data relating to its performance, such as its efficiency, input and output temperatures of thc service fluid, settings and duty cycles of the expansion valve and compressor or temperatures and pressures of the refrigerant at one or more points in the closed loop.

Diagnostic systems within the heat pump may analyse this information to detect sub-optimal operation of the system or faults within the heat pump and may notify the user or a central control system. Alternatively, raw data may be transmitted to a control system for remote analysis and diagnostics.

In particular, data derived from the heat pump may be used to alert a user or service provider to faults in the heat pump and to advise how such faults might be addressed, for example by adjusting refrigerant levels or components of the refrigerant or by replacing parts of the system.

Aspects of the system described above may be implemented independently or in conjunction with one another. They may be implemented in systems designed for heating or for cooling a target environment, as described above.

Embodiments of the system and method will now be described in more detail with reference to the figures in which: Fig. lisa schematic diagram of a heat pump; Fig. 2 show the dew point and bubble point lines for a zeotropic mixture of ammonia and water; Fig. 3 is a graphical representation of the relationship between the critical temperature and the number of carbon atoms in an alkane; Fig. 4 is a graphical representation of the reverse Carnot refrigeration cycle; Fig. 5 is a simplified representation of the reverse Camot refrigeration cycle; S Fig. 6 is a schematic diagram of a heat pump having a suetionlliquid heat exchanger; Before setting out the details of the proposed heat pump system and refrigerant, further

background information may be helpful.

In the UK, after allowing for transmission loses, and less efficient power stations, the overall efficiency of the national grid is now better than 40%. Therefore heat pumps with Seasonal Co-efficients of Performance (SCOPs) greater than 2.5 will produce renewable energy. A maximum co-efficient of performance for heating may be calculated using the formula below, where the temperatures are expressed in units of Kelvin (K): Thot COPbeatan = Thot-Tcool For example for a GSHP with a condensing temperature of 60°C (333K) based on a heat source at 5°C, the theoretical COP heating is 6. However, such performances are not achieved in practice.

For a typical ground source heat pump with ground loop at -3-0°C, evaporating at -9°C and hot water at 35 / 29°C, condensing at 41°C, the theoretical Carnot COP is 6.28, but in reality only 3.37 is achieved, hence the Carnot efficiency is referred to as being 3.37 x 100 / 6.28 = 53.7%.

For a high efficiency ground source heat pump with ground water at II-6°C, evaporating at 0°C and hot water at 35 / 29°C, condensing at 41°C, the Carnot COP is 7.66, but in reality only 4.75 is achieved, hence the Carnot efficiency is referred to as being 4.75 x 100 / 7.66 = 62.0% In embodiments described herein, because of the temperature glide, the compressor "sees" an evaporating temperature of 2°C and a condensing temperature of 72°C, so the equivalent Carnot COP is 4.93, so in order to achieve an actual COP of 3.37, we need to achieve a Carnot efficiency of 68.4%, only 6.4% better than existing high efficiency heat pumps.

Embodiments of the present system may achieve this because of at least one of the factors of: the use of much higher critical temperatures, the use of temperature glide, the use of a suction liquid hcat exchanger, and/or the use of hydrocarbon refrigerants.

Recent increases in fossil fuel prices, together with improving efficiencies from heat pumps mcan that heat pumps now provide a heating option that is financially highly competitive.

However, capital costs are higher for heat pump systems and when considering the purchase of a heat pump for economic reasons pay back periods relative to mains gas heating can be long. In view of the high capital costs, most heat pump systems cannot compete economically with mains gas in the UK, but they provide a competitive option when compared to off-peak electric heating, LPG and oil-based heating.

There are over one million households in the UK without a mains gas supply for whom the current main heating options are off-peak electric, LPG and oil-based heating. Since most homes with LPG and oil are in rural locations, the ahemative of mains gas supply is not economically feasible.

Home and property owners arc seeking to reduce their heating energy costs and the governments around the world are seeking to reduce significantly their CO2 emissions.

Heat Pumps are expected to make a major contribution to both these objectives.

Current heat pumps have the shortcoming of the need to retrofit under floor heating at great cost and inconvenience to the home/property owner. Therefore, a mass market for heat pumps will only emerge when they can be retrofitted to existing wet radiator heating systems in the existing domestic housing stock.

Traditionally Low Temperature Hot Water (LTHW) heating systems have been designed at 82°C flow / 71°C return to give an average water temperature of 76.5°C. Since 1997 the European standard EN442 reduced temperatures to 75°C flow / 65°C return, giving an average of 70°C. However, this is still too hot for a condensing boiler to operate in condensing mode which requires 70°C flow / 50°C return; producing 26.5% less heat output from EN442 designed radiators. Moreovcr, typical hcat pumps operatc at 50°C flow / 42°C return, producing 57.7% less heat output from EN442 designed radiators.

Furthermore, neither condensing boilers operating in condensing mode nor typical heat pumps can raise stored domestic hot water to a safe temperature in order to prevent legionella growth.

Using the 1-lydrofluorocarbon (HEC) refrigerant R-l34a in economised scroll compressors it may be possible to achieve water temperatures up to 70°C flow / 60°C return, giving an average of 65°C but this still results in a significant drop in hcat output from EN442 designed radiators.

Environmental motives for purchasing heat pumps are compromised when using HFC refrigerants which are restricted under the Kyoto Protocol due to their very high global warming potential, more than 1000 times higher than CO2. HFCs cause significant environmental impact during their manufacture, as a result of leakage during use, and during disposal at the end of the equipment life. Therefore, such refrigerants are not considered to be environmentally friendly and the relative impact of this aspect increases significantly as electricity generation is decarbonised to such an extent that if electricity generation were to be fitlly decarbonised, then 100% of the environmental impact of a heat pump would be derived from the environmental properties of the refrigerants.

It would therefore be advantageous to develop an environmentally friendly refrigerant with low global warming potential and the ability to deliver an average water temperature of 70°C without detriment to the SCOP. Further advantages would arise if the same refrigerant could be used for air conditioning in very high ambient regions.

One embodiment of a heat pump 100 is illustrated schematically in Fig. 1. The heat pump 100 includes a closed loop 110 around which circulates a refrigerant fluid. The heat pump also includes a compressor 112 to drive the refrigerant fluid around the closed loop 110.

The effect of the compressor 112 is to raise the pressure of the refrigerant fluid in one half of the closed loop I lOa. The fluid passes from the compressor 112 around the closed loop until it reaches the expansion valve 114. As the fluid passes through the expansion valve 114, it expands and cools. It then continues around the closed loop 110 back to the compressor 112.

Having passed through the expansion valve 114, cooled refrigerant fluid passes through a first heat exchange zone or evaporator 116 in which heat is absorbed into the refrigerant fluid from an external heat source 120. This is usually achieved by passing a fluid, such as water, through pipes running through the heat source, which would be the ground in the case of a ground source heat pump. The pipes carrying the fluid from the heat source 120 are then incorporated into a heat exchanger system together with the closed loop carrying the refrigerant in the evaporator 116.

Heat absorbed from the heat source 120 evaporates the refrigerant so that it enters the compressor 112 in a gaseous form.

The gaseous refrigerant that exits the compressor 112 at a relatively high temperature and pressure passes through a second heat exchange zone or condenser 118 in which heat is transferred from the refrigerant to a heat sink 122, such as a domestic or commercial property. Again, the heat in the refrigerant is usually transferred to the heat sink via a second heat exchanger 118 which transfers the heat to a further fluid such as water circulating in the heating system of the heat sink 122. This transfer of heat cools the refrigerant fluid in the closed loop 110, which is then cooled further by expansion through the expansion valve 114, as set out above.

Some of the terminology used herein will now be described in more detail for ease of reference. However, it is noted that the descriptions below are provided as background information only and are not intended to be limiting in any way.

The critical temperature of a substance, which may also be referred to as its critical point or critical state, can be considered to be the temperature at which, for a given pressure, the phase boundaries of the substance cease to exist. At or above the critical temperature, the substance no longer exists in a distinct liquid or gas phase but the properties of its gas and liquid phases have converged and the substance forms a supercritieal fluid.

When an azeotropic mixture of substances reaches its boiling point, the resulting vapour has thc same proportions of components as the original mixture. Therefore, thc proportions of components in an azeotropic mixture cannot be altered by distillation. At at least one pressure, the boiling points of the components of the mixture are essentially the same.

In a zeotropic mixture, however, the component substances have different boiling points so, as the mixture reaches the temperature that is the boiling point of the lowest-boiling-tcmperaturc component in the mixture, which is known as the bubble point of the mixture, a first substance boils and is distilled out of the mixture leaving the remaining components in a liquid state. Therefore, the proportionate constituents of the zcotropic mixture change as the mixture is heated beyond the bubble point to the temperature of the substance with the highest boiling point, above which the whole mixture then exists in a gaseous phase.

This means that components of the mixture are distilled out of the mixture as the temperature is raised above the bubble point. Similarly, there is a difference in the temperatures of condensation of component substances in the mixture. The highest temperature at which one component of the mixture condenses is the dew point of the mixture. Therefore, taken as a whole, there is a temperature difference between the bubble point and the dew point of a zeotropic mixture.

The glide of a zeotropic mixture is the difference, at a particular pressure, between the bubble point temperature and the dew point temperature of the mixture. That is, the difference between the minimum boiling point of a component in the mixture and the maximum condensation point of a component in the mixture.

Fig. 2 is a diagram illustrating the bubble and dew points for different zeotropie mixtures of ammonia and water at a constant pressure of 16 bar. The bubble and dew points for the mixture are different except at the points where the mixture comprises 100% water or 100% ammonia.

It is noted that a particular mixture of substances can be either azeotropic or zeotropic depending on the pressure at which it is held.

It has been appreciated that a refrigerant suitable for use in a new GSHP should advantageously have a critical temperature higher than that of other refrigerants such as R- 134a. It is advantagcous to dcsign a rcfrigcrant with a critical temperaturc wcll above the required condensing temperature since thermodynamic properties and principles result in declining efficiency as refrigerant condensing temperatures increase and approach the cntical temperature.

In addition, it has been appreciated that, for practical, compactness and equipment cost reasons, it can be adyantageous to minimise the normal boiling point of the refrigerant to give the highest pressure. Unfortunately, there is a negative correlation between the normal boiling point and the critical tcmperaturc; increasing critical temperaturc incvitably results in decreasing normal boiling point.

The problem can be seen even more graphically when studying unsaturated hydrocarbon rcfrigerants whcrc it bccomes apparcnt that it is ncccssary to maximisc the number of carbon atoms in the molecule in order to maximise the critical temperature, which inevitably raises the boiling point. Fig. 3 shows the critical temperature of some alkanes as the number of carbon atoms in the alkane is increased. Therefore, we need to optimise the balance between critical temperature and normal boiling point.

To do this, we can consider the Carnot cycle. Fig. 4 illustrates the classic reverse Carnot refrigeration cycle, where TN is the condensing temperature, T is the evaporating tcmpcraturc and the peak of thc graph is at thc critical temperature.

Fig. 4 can be simplified further to a perfect square-shaped Carnot Cycle as illustrated in Fig. 5, with a scmi-circlc for thc dome. It can thcn bc sccn using gcometry that thc critical temperathre shoidd be higher than the condensing temperature by half the difference between the condensing and evaporating temperatures since, for a semi-circle, D=2R.

However, real refrigerants in real equipment do not follow the reverse Carnot cycle and analysis of empirical data from cooling equipment operating in high ambient conditions, suggests that in practice, the critical temperature must be higher than the condensing temperature by 100% of the difference between the condensing and evaporating temperatures.

For example in a heat pump design with a water-off temperature of 75°C, we might select a S condensing temperature of 85°C, and if ground coupled at 11°C, an evaporating temperature of 5°C. The difference between the evaporating and condensing temperatures is 80 and if we add this figure to the condensing temperature we get 165. This suggests that we should be looking for critical temperatures in excess of 165°C for optimum efficiency and we can use this benchmark to assess potential replacements for heat pump and high temperature cooling applications. So to achieve this in a blend we can consider potential constituents with a critical tcmpcraturc abovc 150 °C and normal boiling point bclow 15°C.

After excluding highly toxic substances, ozone depleting substances (those containing chlorine, bromine and iodine) and high global warming substances (those with more than one fluorine atom), 18 potential candidates arc proposed for suitable components for a refrigerant. These are listed below in order of critical temperature.

Critical Normal Temperature, Boiling Name of Potential Refrigerant °C Point, °C 1-Butyne (Ethylacetylene) 190.5 8.0 cyclobutanc 186.8 12.5 1,2-Butadiene (Methylallene) 176.0 10.8 Cyclobutene 173.2 2.0 Methoxyethene ( Methyl vinyl ether) 172.0 6.0 Methoxyethane ( ethyl methyl ether) 164.7 7.4 Dimethylamine 164.6 7.0 Methylcyclopropane 164.0 0.7 Cis-2-butene 162.5 (174.9) 3.7 neopentane 160.6 9.5 trimethylamine 160.1 2.9 sulphur dioxide 157.6 -10.0 Fluoropropene 157.34 -1.7 Methylamine 156.9 -6.0 Trans-2-butcnc 155.5 (166.3) 0.9 butane 152.1 -0.5 1,3-Butadiene 152 -4.5 Disilane 150.8 -14.5 Any of the above refrigerants may be suitable for use in a new refrigerant blend as described herein. However, this list of candidates can be further reduced by selecting stable, fluorine-free, non-toxic refrigerants and removing potential refrigerants with a higher normal boiling point than the one above.

Name of Potential Critical Temperature, Normal Boiling Refrigerant Point, °C cyclobutane 186.8 12.5 1,2-Butadicnc (Mcthylallcnc) 176.0 10.8 Cyclobutene 173.2 2.0 Trans-2-butene 155.5 (166.3) 0.9 Butane 152.1 -0.5 I,3-Butadiene 152 -4.5 Disilane 150.8 -14.5 This gives a list of seven refrigerants from which to select an efficient refrigerant. In the previous example of a heat pump design with a condensing temperature of 85°C and an evaporating temperature of 5°C and a target critical temperature in excess of 165°C, Cyclobutene is the optimum single component refrigerant. However, the above list of refrigerants is not optimised in terms of what can be achieved by looking at blends.

Onc option exists in the form of zeotropic blends. As set out above, zcotropcs have a difference between the bubble point and dew point due to the different boiling points of the constituents. This difference is called the glide value for the refrigerant. Zeotropic designs uses temperature glide to advantage, by adopting counter current evaporator and condenser heat exchangers which allow co-current flow of the refrigerant liquid and vapour mixtures, as described in more detail below. By matching the refrigerant side temperature glide with the service fluid temperature change, counter flow design lowers discharge pressures and raises suction pressures, so that the compressor does less work to deliver the same amount of heating (or cooling). Power savings of up to 25% can be achieved relative to comparable single component refrigerants.

One proposed blend will be designed to condense between 80°C and 72°C to produce hot water at 74°C flow and 66°C return, giving an average of 70°C in compliance with EN442.

On the cold side of the heat pump, the refrigerant will evaporate between 108°C and 2°C whilst cooling a CO2 thermosyphon or ground water from 12°C to 4°C. The resultant S SCOP has been estimated at 3.7 which is midway between the typical ground source heat pump and the best ground source heat pump figures even when coupled to under floor heating and well above the threshold figure of 2.5 for the production of renewable energy.

In one embodiment, the blend or mixture comprises I,2-Butadiene (R2380a) / Cyelobutene (50.0 / 50.0). 1-lowever, the relative proportions of these components may vary and further components may be added to increase the tcmperature glidc or to vary the prcssure. For example, the addition of cyclobutane (RC390) would lower the pressure, whereas the addition of Trans-2-butene, butane (R600), 1,3-Butadiene, or Disilane would raise the pressure.

Fig. 6 is a schematic diagram of a ground source heat pump according to one embodiment.

The GSHP of Fig. 6 includes a compressor 610, a condenser 612, an expansion valve 614 and an evaporator 616 that operate according to the same principles as set out above with reference to Fig. 1. As illustrated in Fig. 6, energy QEvAP is transferred from the heat source, in this case the ground, to the closed refrigerant loop 618 in the evaporator 616 and energy QUOND is transferred from the closed refrigerant loop 618 to the target environment in the condenser 612.

In addition, the GSHP of Fig. 6 includes a Suction Line Heat Exchanger or Suction Liquid Heat Exchanger (SLHX) 620, which enables exchange of heat between refrigerant in the closed loop 618 that is about to enter the compressor 610 and refrigerant that has passed through the condenser 612. The refrigerant that has passed through the condenser 612 has already transferred some of its energy to the target environment, but is still at a relatively high temperature and pressure when compared to the refrigerant that has not yet entered the compressor. Therefore, energy is transferred in the SLHX to the lower-temperature refrigerant, raising the temperature and pressure of this refrigerant before it enters the compressor 610. This means that the compressor 610 has less work to do in raising the refrigerant to the temperature and pressure necessary for it to enter the condenser 612 and therefore effectively increases the efficiency of the compressor 610. This contributes to increased evaporator performance and increased heat extraction from the energy source.

The SLHX can also act to increase the lifetime of the compressor.

Use of the SLHX can be particularly advantageous for a zeotropic refrigerant because it reduces or eliminates the risk of liquid carryover from the evaporator to the compressor.

Any liquid carryover could dilute the compressor oil, risking compressor failure and zeotropic blends are more prone to liquid carryover because the more volatile component can entrain liquid droplets of the less volatile component.

In prefcrrcd embodiments, the cxpansion valvc 614 has an associated controller (not shown) that controls the expansion of the refrigerant to ensure it exits the expansion valve at a desired pressure and temperature. The pressure may be controlled to within a range of acceptable pressures, or the valve may be controlled to expand the refrigerant to an optimum pressure for the system. In some embodiments, the desired pressure may be varied depending on factors including the desired temperature for the target environment and/or the current temperature of the heat source or the target environment. Preferably, the expansion valve may be controlled based on a temperature sensor located between the SLHX 620 and the compressor 610. This allows the evaporator to operate without superheat, in other words all of the heat exchange surface area can be used for evaporation, i.e. converting liquid into gas, rather than heating the gas above its evaporating temperature (or in the case of a zeotropic refrigerant its dew point).

The compressor 610 itself may be a belt-driven compressor with a power from 1kW to 6kW for domestic applications, or preferably a hermetic scroll compressor, for example similar to the types used to compress mains gas. Alternative compressor designs will also be investigated including types more commonly used for the compression of air.

The condenser 612 operates to transfer heat between the refrigerant and the target environment via a service fluid, typically water, that flows through a heat exchanger in the refrigerant. Typically, the service fluid enters the condenser at a temperature of around 65°C and exits the condenser at a temperature of 75°C. This increase in temperature within the condenser is preferably matched to the drop in temperature of refrigerant and preferably is also matched to the temperature glide in the zeotropic mixture of the refrigerant.

Typically, the refrigerant enters the condenser at a temperature of around 80°C and exits the condenser at a temperature of around 72°C. The temperature at which the service fluid enters and preferably exits the condenser can be used to control the operation, in particular the capacity, of the compressor.

In one specific, non-limiting embodiment, Hydrocarbon Refrigerant Ground Source Heat Pumps described herein may provide a robust, reliable and efficient technical solution which maintain a very high COP, including integrated delivery of domestic hot water at 74°C. Some designs make use of inverter controls to improve the compressor lifetime by reducing short cycling. The use of a refrigerant suction! liquid exchanger can increase the life expectancy of the heat pump and increase efficiency. This can contribute to increased evaporator performance and increased heat extraction from the energy source.

Product Details of a specific embodiment may include one or more of the following features: * Compact design and robust construction for tidy installation and low noise levels.

* Preset temperature and easy to use automatic controls.

* Very easy to use.

* All units are pre-charged with hydrocarbon refrigerant at the factory.

Hydrocarbon coolant is efficient, natural and environment friendly.

* The use of hydrocarbon refrigerant reduces wear and tear on the compressor thus increasing its life span.

* A built in automatic safety system protects the heat pump from overloading.

* The heat pump is delivered complete with electrical fittings including automatic control and plugs.

* Flexible and easy to change parts.

In one embodiment, the system comprises a Geothermal Heating Boiler (GeoBoiler) that is suited to retrofit or new property build, where conventional radiators are used as the space heating system and indirect tanks are used for hot water. The boiler can be used as a direct replacement for current oil, LPG and natural gas boiler systems.

Embodiments of the heat pump descried herein may use borehole or probe heat sources, with preference to pumped / thermo-siphon C02 probes since these offer greater overall efficiency. However, the system may also be implemented with water-based heat source systems. in embodiments that operate using C02-based systems, particularly if the heat pump is to be installed internally in a domestic environment, the C02 may be odourised or mixed with another gas canying an odour to enable users to detect any leakage of the asphyxiating C02 from the system.

In some embodiments, a further controller may be provided to turn the heat pump on or off.

The controller may bc operable remotely, for example from a central control centre or from a service provider. In some embodiments, the controller may operate automatically. This may be provided as a safety feature, to prevent operation of the system if it becomes dangerous or faulty. It may also enable a service provider to have control over the system.

Alternatively, or in addition, the heat pump could be used to provide high or low frequency response to a power grid in the form of dynamic demand by turning the heat pump on or off to extract energy from the grid in times of over-supply of electrical energy on the grid or to desist from using energy from the grid in times of over-demand. Operation of such a system may be controlled automatically, for example by monitoring for frequency changes on the power grid. In such an embodiment, the heat pump may be turned on or off only on the condition that the temperature of the target environment remains within a predetermined temperature range. This ensures that the frequency response feature does not detract from the primary heating or cooling function of the heat pump.

Claims (27)

1. Claims 1. A refrigerant comprising a plurality of molecules forming a zeotropic mixture having a high critical temperature and a wide glide comprising a wide temperature difference between the temperature at the bubble point and the temperature at the dew point of the mixture at a particular pressure.
2. 2. A refrigerant according to claim I wherein the temperature difference between the bubble point and the dew point of the mixture is at least 7°C, preferably at least 8°C.
3. 3. A refrigerant according to any preceding claim wherein the refrigerant has a critical temperature that is higher than the dew point temperature by at least half the difference between the dew point and bubble point temperatures, preferably by at least the difference between the dew point and bubble point temperatures.
4. 4. A refrigerant according to claim I or 2 wherein the refrigerant has a critical temperature greater than 135°C, preferably greater than 150°C, frirther preferably greater than 165°C.
5. 5. A refrigerant according to any preceding claim wherein at least one molecule, preferably all molecules, has a critical temperature greater than 135°C, preferably greater than 150°C, ffirthcr prcfcrably grcatcr than 165°C.
6. 6. A refrigerant according to any preceding claim wherein the temperature at the dew point of the mixture at atmospheric pressure is less than 15°C, preferably less than 12°C.
7. 7. A refrigerant according to any preceding claim wherein the temperature at the bubble point of the mixture at atmospheric pressure is greater than 0°C.
8. 8. A refrigerant according to any preceding claim wherein the temperature at the dew point of the mixture at a pressure of 300psi is greater than 65°C, preferably greater than 70°C.
9. 9. A refrigerant according to any preceding claim wherein the plurality of molecules comprise hydrocarbon moiccuics.
10. 10. A refrigerant according to any preceding claim wherein the molecules each comprise one or zero fluorine atoms.
11. 11. A refrigerant according to any preceding claim, the refrigerant comprising at least one molecule selected from a first group consisting of: cyclobutane and 1,2-butadiene.
12. 12. A refrigerant according to any preceding claim, the refrigerant comprising at least OIIC molecule selected from a second group consistillg of: cyclobutene, trans-2-butene, butane, 1,3-butadiene, disilalle and fluoropropene.
13. 13. A rcfrigcrant according to claim 12 as dcpcndcnt on claim 11, whercin at least 80%, prcfcrably at Icast 90% by weight of thc rcfrigcrant comprises a combination of molecules selected from the first and second group.
14. 14. A refrigerant according to claim 12 as dependent on claim 11, wherein the refrigerant comprises a combination of at least 25% by weight of a molecule selected from the first group of molecules and at least 25% by weight of a molecule selected from the second group of molecules.
15. 15. A refrigerant according to any prcccding claim, comprising 40-60% by wcight of 1,2-butadielle and 60-40% by weight of cyclobutene.
16. 16. A refrigerant according to any preceding claim comprising substantially equal proportions by weight of 1,2-butadicnc and cyclobutenc.
17. 17. A refrigerant comprising a plurality of molecules forming a zeotropic mixture having a critical temperature greater than 135°C and a wide glide comprising a temperaturc difference at lcast 7°C between the tcmperaturc at the bubble point and the temperature at the dew point of the mixture at a particular pressure.
18. 18. A heat pump comprising: a ground-coupled evaporator; and a rcfrigcrant according to any of claims 1 to 17.
19. 19. A heat pump comprising: a ground coupled evaporator; a hydrocarbon refrigerant; the heat pump having a nominal average working temperature of at least 65°C.
20. 20. A heat pump according to claim 19, wherein the nominal average working temperature is 70°C.
21. 21. A heat pump according to any of claims 18-20 further comprising means for mixing the refrigerant.
22. 22. A heat pump according to any of claims 18-21 wherein the nominal output temperature of the heat pump is greater than 65°C, preferably greater than 70°C.
23. 23. A heat pump according to any of claims 18-22 further comprising pressure reduction means for reducing the temperature and pressure of the refrigerant and control means for controlling the operation of the pressure reduction means.
24. 24. A method of determining the composition of a refrigerant in a heat pump, the method comprising obtaining a measure of the thermal characteristics and/or performance of the heat pump and analysing the measure to determine the relative proportions of components in the refrigerant of the heat pump.
25. 25. A method according to claim 24 including providing an indication of a component of the refrigerant that is higher than a desired level.
26. 26. A refrigerant substantially as described herein.
27. 27. A heat pump being substantially as described herein with reference to and illustrated by Fig. I or Fig. 6.