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Written evidence submitted by Nicholas Cox, Director of Sustainable Working Fluids Ltd

Heat pumps extract energy from the ambient environment and use electricity to upgrade this energy to a useful temperature. After allowing for transmission loses, and less efficient power stations, the overall efficiency of the national grid is now better than 40%. Therefore heat pump seasonal COPs greater than 2.5 will produce renewable energy. Recent increases in fossil fuel prices, together with improving efficiencies from heat pumps mean that for the first time since the early 1980s higher efficiency heat pumps are once again the cheapest source of heating in the UK as shown in Table 1:

Heating System Price CO2 emissions Efficiency CO2 Price Annual P/kWh per kWh fuel % emissions per running

(Kg CO2/kWh) per kWh useful cost for 4 useful heat heat bed (Kg output detached

CO2/kWh) p/kWh house LPG Condensing Boiler 10.67 0.25 90 0.28 11.86 1690 Off peak electric 6.27 0.41 80 0.51 5.99 1117 Oil boiler (28 sec) 6.0 0.27 80 0.34 6.95 1069 Coal boiler 3.90 0.29 70 0.41 5.57 794 Biomass (Wood pellets) 5.16 0.03 70 0.04 7.37 1,050 Mains gas Condensing boiler 4.90 0.19 90 0.21 5.44 776 Typical air source heat Pump 10.431 0.41 220 0.12 4.74 676 High efficiency air source heat 10.43 0.41 310 0.11 3.36 478 pump Typical ground source heat 10.43 0.41 337 0.11 3.09 440 pump High efficiency ground source 10.43 0.41 475 0.08 2.20 313 heat pump

Table 1: Comparison of Different Heating Systems2

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. Maximum market values for heat pumps have been calculated in Table 2.

Heating System LPG Off peak Oil Coal Biomass Mains gas electric Typical air source heat Pump 10,410 4410 3930 1180 3740 1000 High efficiency air source heat 12,120 6390 5910 3160 5720 2980 pump Typical ground source heat 12,500 6770 6290 3540 6100 3360 pump High efficiency ground source 13,770 8040 7560 4810 7370 4630 heat pump

Table 2: Maximum additional investment (£s) for 10 year payback period

1 assumes 30% off peak rate usage 6.27 p / 70% peak usage 12.21p 2 Updated by Google search August 2014

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Clearly, heat pumps cannot compete economically with mains gas and people do not buy LPG heating for rational economic reasons. If we assume that people will avoid the solid fuel options for reasons of convenience, we are left with two competing alternatives, off peak electric and oil. With the assistance of grant funding, both high efficiency air source heat pumps and high efficiency ground source heat pump are economically viable against both off peak electric and oil.

There are over one million households in the UK without a mains gas supply, moreover the Government wants 3.8 million new homes built by 2021. The Building Research Establishment (BRE) Domestic Energy Fact File Report 2008 shows in the UK there are 23 million centrally heated dwellings in 2006 of which 1 million are oil-fired and 20 million are gas-fired. It is difficult to predict how future oils and gas process will change, but it seems reasonable to assume that in the case of oil, prices will continue to climb, making oil central heating ownership a less and less attractive proposition. Since most homes with oil are in rural locations, the alternative of gas supply is not economically feasible.

Home and property owners are seeking to reduce their heating energy costs and the UK is seeking to significantly reduce its CO2 emissions by 2020, and further to 2050. Heat Pumps are expected to make a major contribution to both these objectives. Market growth is predicted due to the introduction of the Renewable Heating Incentive (RHI) for GSHP installations. The payments received by a homeowner with a new installation are expected to cover about 75% of the upfront installation cost, making the residual cost the equivalent of a simple oil or gas boiler replacement. Thus the RHI will in effect eliminate, or at a minimum substantially mitigate, the upfront-cost barrier which currently limits the heat pump market in the UK.

The size of the expected GSHP market in 2020 is set out in The UK Supply Curve for Renewable Heat, a Study for the Department of Energy and Climate Change – July 2009, authored by NERA Economic Consulting and the AEA. See particularly Appendix C.2.4.3 - Projections of future deployment. This report details ‘Stretch’ and ‘High’ growth scenarios, but using the more conservative Central Growth Scenario the forecast for 2020 is: 290,000 Domestic units producing 3.7 TWh heat output, and 32,000 Non-domestic units producing 11 TWh heat output.

For 2050, DECCs stated aim is the complete de-carbonisation of domestic heating. At this time domestic gas and oil boilers are expected to have been phased out and been replaced with electricity powered forms of heating (such as GSHPs, Air Source Heat Pumps, and direct thermal conversion) and other zero-carbon renewable sources.

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 LTHW heating systems have been designed at 82˚C flow / 71˚C return to give an average 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, giving an average of 60˚C; producing 26.5% less heat output from EN442 designed radiators. Moreover, typical heat pumps operate PMP0002 at 50°C flow / 42°C return, giving an average of 46˚C; 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 HFC R-134a in economised scroll compressors it is 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 heat output from EN442 designed radiators. Daikin have produced a two stage cascade heat pump, capable of delivering water temperatures of up to 80˚C. In both these instances, the SCOP drops below 2.5, thus becoming uneconomic and failing to produce any net renewable energy.

Environmental motives for purchasing heat pumps are compromised when using HFC which are restricted under the Kyoto Protocol and F-gas Directives, due to their very high , 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. It is unlikely that the government will permit a “new” large scale market for HFCs given that the switch from sunroofs to HFC air conditioning in cars has played such havoc with global warming projection figures. At the very least, current heat pump grants for HFC containing equipment must be short lived.

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. This requires a refrigerant with a critical temperature significantly higher than the critical temperature of R-134a. There is exactly the same requirement for air conditioning in very high ambient regions. In May 2010 the Montreal Protocol Technology and Economic Assessment Panel (TEAP) “Scoping Study on HCFC Alternatives under High Ambient Temperature Conditions” was published. This summed up the problem perfectly:

“The critical temperature of a refrigerant is an important parameter in the effectiveness and efficiency of equipment unless explicitly designed for transcritical operation as is typical for systems designed to use R-744 () as the refrigerant. In the conventional vapour-compression cycle based equipment the condensing temperature is kept well below the critical temperature, because thermodynamic properties and principles result in declining capacity and efficiency as heat-rejection (refrigerant condensing) temperatures increase and approach the critical temperature.”

The following table lists 30 industry recognised zero ODP refrigerants in order of critical temperature. Surprisingly perhaps, HFC R134a only gets half way up the league table – fourteen refrigerants have a higher critical temperature and hence higher potential for energy efficient operation. PMP0002

Refrigerant number Refrigerant Name Critical Temperature, °C

1 R718 Water 374.0 2 R601a 2-methylbutane () 187.3 3 RC390 cyclobutane 186.8 4 R1336mzz hexafluorobut-2-ene 171.3 5 R764 sulphur dioxide 157.6 6 R630 156.9 7 R245fa 1,1,1,3,3- 154.05 8 R600 n- 152.1 9 R600a 134.7 10 R717 132.3 11 R510A RE170/R600a (88.0/12.0) 128.1 12 RE170 Methoxymethane () 127.2 13 R436B R290/R600a (52.0/48.0) 115.0 14 R1234ze 1,3,3,3-tetrafluoropropene 109.4 15 R134a 1,1,1,2-tetrafluoroethane 101.1 16 R432A R1270/E170 (80.0/20.0) 99.4 17 R290 96.8 18 R1234yf Tetrafluorprop-1-ene 95.0 19 ECP410A R1270/R161 (75.0/25.0) 94.9 20 R1270 (propylene) 92.4 21 R407C R-32/125/134a (23.0/25.0/52.0) 86.0 22 R32 78.1 23 R404A R-125/143a/134a (44.0/52.0/4.0) 72.0 24 R410A R-32/125 (50.0/50.0) 71.4 25 ECP717 R170/R717 (45.0/55.0) 41.9 26 R422B R-125/134a/600a (55.0/42.0/3.0) 41.3 27 ECP744 R744/R41 (50.0/50.0) 37.9 28 R744A 36.4 29 R170 35.0 30 R744 carbon dioxide 31.0

Table 3: Thirty zero ODP refrigerants in order of critical temperature

Based on the above table, it would be reasonable to assume that R718 (water) is the perfect refrigerant. So why hasn’t R718 been adopted as the universal refrigerant? The problem is the very large size and high cost of the equipment required for a refrigerant with such a very low volumetric capacity, high normal (100˚c!), high specific volume, high pressure ratio, and resulting high compressor outlet temperatures. The technology was first showcased in the Lego factory in Denmark in the mid 1990s but all subsequent attempts at fully commercial production have failed. However, a recent internet platform “Everything R718“ 3 established as the industry platform for the water used in cooling and refrigeration, suggests that commercialisation remains a possibility.

In practice we have two conflicting commercial pulls - for efficiency, critical temperature is the most important factor, whereas for practical, compactness and equipment cost

3 http://www.r718.com/ PMP0002

reasons we want the lowest normal boiling point to give the highest pressure. The graph below demonstrates the negative straight line correlation between the two, increasing critical temperature inevitably results in decreasing normal boiling point.

Refrigerants with a critical temperature above R134a (100.9 °C) and normal boiling point below R123 (27.8°C)

250 C °

200

, e r u t

a 150 r e p m

e 100 T

l a c i t i

r 50 C

0 -40 -30 -20 -10 0 10 20 30 40 Normal Boiling Point, °C

Figure 1: correlation between critical temperature and normal boiling point.

The problem can be seen even more graphically when studying unsaturated refrigerants where it becomes apparent that it is necessary to maximise the number of carbon atoms in the molecule in order to maximise the critical temperature4, which inevitably raises the boiling point.

4 Thermophysical Properties of Chemicals and , Carl L. Yaws PMP0002

Figure 2: Critical Temperature of Normal Alkanes

We therefore need to find a way to optimise the balance between critical temperature and normal boiling point. Traditionally this process has been deliberately complicated by refrigerant manufacturers to preserve exclusivity in the refrigerant developers’ “club” and to limit choice in the wider market. This conspiracy against the laity can be over come by a back to basics approach. Firstly, we need to go back to the basic reverse Carnot refrigeration cycle:

Figure 3: Classic Reverse Carnot Refrigeration Cycle

If the above is simplified further, to a perfect square shaped Carnot Cycle, with a semi- circle for the dome, then using basic schoolboy geometry, it becomes obvious that the critical temperature must be higher than the condensing temperature by half the difference between the condensing and evaporating temperatures, because for a semi-circle, D = 2R. PMP0002

Figure 4: Simplified Carnot Cycle

Unfortunately 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, particularly in the Australian mining industry, 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 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 temperature above 150 °C and normal boiling point below 15°C.

The chemical industry keeps telling us that we are running out of potential refrigerants. So it is surprising that, even 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), there are some 18 potential candidates5,6,7.They are listed in order of critical temperature as this is the best initial indication of their likely efficiency as refrigerants:

Critical Normal Refrigerant Temperature, Boiling number Refrigerant Name °C Point, °C 5 Recommended Critical Temperatures. Part I. Aliphatic Hydrocarbons, Iwona Owczareka… and Krystyna Blazejb, InstituteUnstable? of Coal Chemistry, Polish Academy1-Butyne of Sciences, (Ethylacetylene) 44-100 Gliwice, Poland; published190.5 4 August 2003 8.0 6 ThermophysicalRC390 Properties of Chemicals and Hydrocarbonscyclobutane by Carl L. Yaws (13 Aug 2008)186.8 12.5 7 http://encyclopedia.airliquide.com/encyclopedia.asp PMP0002

R2380a 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 Unstable? Methylcyclopropane 164.0 0.7 Cis-2-butene 162.5 (174.9) 3.7 R601b neopentane 160.6 9.5 trimethylamine 160.1 2.9 R764 sulphur dioxide 157.6 -10.0 Fluoropropene 157.34 -1.7 R630 Methylamine 156.9 -6.0 R1390 Trans-2-butene 155.5 (166.3) 0.9 R600 butane 152.1 -0.5 R2380 1,3-Butadiene 152 -4.5 Disilane 150.8 -14.5

Table 4 – Potential refrigerants with a critical temperature above 150 °C and normal boiling point below 15°C.

The first problem is that there are far too many candidates. So for a first iteration, we should consider stable, fluorine free, non toxic refrigerants. The following table excludes any refrigerant with a higher normal boiling point than the one above to produce a shortlist of desirable refrigerants:

Refrigerant Refrigerant Critical Temperature, Normal Boiling number Name °C Point, °C

RC390 cyclobutane 186.8 12.5 R2380a 1,2-Butadiene (Methylallene) 176.0 10.8 Cyclobutene 173.2 2.0 Trans-2-butene 155.5 (166.3) 0.9 R600 butane 152.1 -0.5 1,3-Butadiene 152 -4.5 Disilane 150.8 -14.5

Table 5 - First Choice, stable, fluorine free, non toxic refrigerants

This gives a list of seven refrigerants from which a heat pump designer could select the most efficient refrigerant. In our 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 , clearly 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.

A better option exists in the form of zeotropic blends. Zeotropes 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. 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 PMP0002 savings of up to 25% have been achieved relative to comparable single component refrigerants.

An optimised blend could 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 10.8˚C and 2˚C whilst cooling a CO2 thermosyphon or ground water from 12˚C to 4˚C. The resultant 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.

Refrigerant Refrigerant Critical Temperature, Normal Boiling number Name °C Point, °C

RC390 cyclobutane 186.8 12.5 R2380a 1,2-Butadiene (Methylallene) 176.0 10.8 zeotropic blend 174.6 2.0 to 10.8 Cyclobutene 173.2 2.0 Trans-2-butene 155.5 (166.3) 0.9 R600 butane 152.1 -0.5 1,3-Butadiene 152 -4.5 Disilane 150.8 -14.5

Table 6 – zeotropic blend properties relative to single component alternatives

This approach using a high critical point refrigerant is not in itself innovative and has been applied many times. The innovation is in the application to GSHPs and the development of novel refrigerant blends.

High critical temperature refrigerant working fluids have been used previously in industrial heat pumps (CFCs R11 and R114, the HCFCs R123, R141B, and R124, and the HFCs R245CA, R236FA, R245FA, and wide glide zeotropic mixtures have been proposed previously ( Prof Hazeldene, Leeds University et al) But they have never been applied to domestic heating situations. Moreover the combination of high critical temperature natural working fluids, wide glide zeotropic mixtures, and ground coupled evaporators to maintain above zero evaporating conditions has never been applied to the problem of how to produce an efficient heat pump at domestic radiator conditions.

The first iteration of the blend will be 1,2-Butadiene (R2380a) / Cyclobutene (50.0 / 50.0), but the final iteration will depend on the results of practical testing as well as economic and health & safety considerations.

October 2020