HEATING TECHNOLOGY
pressure in the system. With higher pressures comes the need for better joints to prevent leaks. The higher the pressure, the greater the chances of equipment failure if the equipment is not regularly maintained. The difference between the evaporator pressure and the condenser pressure (plus the amount of super heat) needs to match both the temperatures required from the hot and cold sides. With boiler replacements these are fixed further apart than normal, the hot side at 80 °C, and the cold side at winter ambient temperatures.
n Expansion valve The near vertical line to the left in Figure 1 is the operation of the expansion valve; these components have become much more sophisticated in recent times, regulating the flow through the circuit, and allowing the pressures to be maintained at the target values in the condenser and evaporator. The slight inclination of the line represents the parasitic thermal losses at this point in the circuit.
n The critical point for refrigerants The top of the curve (in Figure 1) is the ‘critical point’ – the pressure above which there is no latent heat of evaporation. Above this there is no liquid to vapour transition. Substance in this region is referred to only as fluid. There is also a specific critical temperature for each substance. Conventional refrigeration circuits are sub- critical, i.e. the entire circuit is below the critical pressure. The curve to the left (in Figure 1) is the saturated liquid condition, where the substance is completely liquid at the boiling point, while the one to the right is the saturated vapour line where the substance is completely vapour at the boiling point. For pure substances and azeotropic blends the transition from saturated liquid to saturated vapour is a horizontal line from the left side to the right. The latent heat of vaporisation is the property that determines the length of the line from the left to the right of the curve. Critical temperatures for some refrigerants used, or capable of operating in the 80 °C condenser range, are shown in Table 1.
n Visualising CoP The CoP for the system is the ratios of the enthalpy for each of the parts of the cycle (accounting for the compressor efficiency). For heating this would be the horizontal length of the red line, divided by the horizontal length of the green line, divided by the compressor efficiency.
REFRIGERANT
CRITICAL TEMPERATURE
R134A (Tetrafluoroethane) 101 °C R290 (Propane)
R32 (Difluoromethane) R407C (Blend) R410A (Blend) R717 (Ammonia)
R744 (Carbon Dioxide)
96 °C 78 °C 86 °C 70 °C
132 °C 31 °C
less onerous external design ambient temperatures. Most of the heat must come from the air. Heat pumps are devices that upgrade the available heat (they expand temperature differences, making the heat more useful); rather than producing the ‘heat’, they simply move it from a cold source to a hot one. All refrigeration cycles (including heat pumps) need
two heat sinks – one ‘cold’ the other ‘hot’. With the cold side, unless there is a low-grade waste heat source from another reliable process, this would normally be the ambient atmosphere. To transfer heat out of the cold source the refrigerant in the evaporator must be at a lower temperature than the cold source temperature (ambient air). This, for each given location, fixes the lower horizontal line in the P-H diagram. To illustrate the source of the heat, Figure 2 shows a
two-stage system. Currently the most likely solution for a boiler replacement heat pump, it illustrates that the number of kW going into the system at the cold side is only slightly less than the number of kW transferred at the hot side. The additional kW coming from the ‘work done’ by the compressors is added to the heat in the system. The CoP of this system would be:
Table 1: Critical temperatures for some refrigerants used, or capable of operating in, the 80 °C condenser range.
Design conditions The external ambient design conditions for a heating system would be more onerous than for a DHW system. Heating systems are likely to need to be able to run 24/7 (in the winter), and the maximum need for heating would be coincident with the coldest ambient. Heating systems would need to work at the design ambient for critical systems (possibly with a margin for climate change). However, the usage of DHW systems is normally low
in the middle of the night, and all heat pump systems for DHW have storage. Thus, it is acceptable to do some of the following: n Delay the recharge until the CoPs are more favourable. n Run the system with a reduced rate of recharge.
DHW systems could be designed for carefully considered,
Transcritical cycles Carbon dioxide has become a common refrigerant for DHS systems, yet its critical temperature is only 31 °C. With a conventional refrigeration cycle the hot side would be limited to something like 25 °C to get a reasonable enthalpy difference across the saturation conditions. This application uses a transcritical cycle, where the evaporator operates in the sub-critical region as normal, and the hot side (no longer actually a condenser, as there is no liquid/vapour transition) operates above the critical point.
Outside the saturation envelope the temperatures
(isotherms) are no longer horizontal, and changing enthalpy at a constant pressure means a related change in temperature. This temperature change is dependent on the water return temperature in the hot side. The problem is that a DHW system is not designed to
lose heat into a process; heat losses at all points in such a system are parasitic. If the water return temperature is too high (say 55 °C in the case of CO2
at 100 bar), then
the expansion of this will still be mostly liquid (at 30 bar) as illustrated in Figure 3 by the central vertical line. As can be seen in Figure 3, the CoP at this point is very low, and possibly less than unity. To get the best CoP out of a CO2
heat pump (in DHW applications), the cold feed needs to provide the lower February 2025 Health Estate Journal 33
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