The control of direct-expansion cooler coils and condensing sets
Remembering that the thermostatic expansion valve does not control cooler coil capacity suitable methods for doing so are as follows.
(a) On-off switching. Simple switching of a single evaporator-condensing set circuit, or of several, independent circuits, may be done satisfactorily from a room or return-air thermostat (together with a step controller if necessary). Typical air change rates in commercially air conditioned premises are 10 to 20 per hour, implying that 6 to 3 minutes is required for the effect of a change in supply air temperature to appear as a change in return air temperature. For a condition of 50 per cent sensible heat gain we would expect equal running and off times for the compressor, suggesting from 12 to 6 minutes between successive starts, which gives 5 to 10 starts per hour under the most adverse partial load condition. The allowable, maximum number of starts per hour should be checked with the supplier of the refrigeration plant. On-off switching must never be attempted from a thermostat located after the cooler because of the certainty of very rapid compressor cycling: when the machine is switched off only a second or two will elapse before the warm, uncooled air that is entering the coil passes over the thermostat and switches the compressor on again.
When several independent cooler coils (each with its own condensing set) are combined for cooling and dehumidification they may be arranged in three ways: in parallel across the face of the air handling unit (to give face control), in series in the direction of airflow (to give row control), or interlaced.
The commonest arrangement is probably when arranged for face control. A return air thermostat switches off the independent coils as a rise in temperature is sensed, allowing uncooled air to by-pass the active coil sections. When they are active, all the refrigeration circuits are equally loaded and this has the great advantage of simplicity and stability. If some of the coil sections are off, the state leaving the coil combination is a mixture of uncooled air that has by-passed the active sections with air that has been cooled and dehumidified.
If the coil sections are arranged in series there is no problem of by-passed air but there are difficulties with the refrigeration circuits, which do not handle equal loads when active. This arrangement is not recommended.
It is possible to arrange coils in series with a common fin block but to interlace their piping circuitry so that each coil section carries the same load when active. A four-row coil, for example, would have a common block of fins, extending over the whole of the cross-sectional area of the air handling unit, but might have two distributors and expansion valves (one for each separate condensing set). Liquid from the distributors would be fed symmetrically through the height, width and depth of the coil, so that each refrigeration piping circuit had the same cooling duty. When one section was switched off, the fins of that section would be available to help cooling and dehumidification throughout the whole of the coil, aiding the performance of the active section and mitigating the by-pass effects suffered by the simple parallel arrangement of face control. As a very simple indication, one circuit might feed rows one and three for the upper part of the coil, with the other circuit feeding rows two and four. For the lower half of the coil, the roles would be reversed, the first circuit feeding rows two and four while the second circuit fed rows one and three. The finning and piping arrangements would be identical for each circuit and the loads would be equalised. In most actual cases the interlacing arrangement is more involved than in the above simplified example.
(b) Cylinder unloading. Since capacity is directly proportional to refrigerant mass flow rate it is also proportional to volumetric flow rate for a given suction pressure. Hence it will be proportional to the number of cylinders in use. The broken line in Figure 12.7 shows the characteristic for the condensing set if half its cylinders are unloaded and air enters the condenser at 10°C. We see now that if the wet-bulb on to the evaporator is 9.6°C the evaporating and suction temperatures will rise to about 3.4°C (point U) and 2.2°C (point V) from their former values of -1.2°C (point Y) and -2.4°C (point Z) respectively. Cylinder unloading can be used to counter the risks of falling suction pressure with reduced wet-bulbs on to the evaporator. See section 12.6.
(c) Back-pressure valve (BPV). If an automatic valve is placed in the suction line (see Figure 12.8) it can be used to give a nominally constant evaporating pressure by increasing the pressure drop in the suction line. Figure 12.8 shows that the drop in the suction line is represented by the points 2 and 3 under the design load with the valve fully open. When the wet-bulb temperature on the evaporator falls the characteristic shifts to the left, balancing at the points 2′ and 3′. By partly closing the valve the natural drop in the suction line (between points 2" and 3") is increased to give a new evaporator balance at the point P. The distance between the points 3" and P represents the pressure drop across the partly closed valve. A pilot valve is sometimes used in conjunction with a thermostatic bulb in the airstream after the cooler coil to give proportional control over the leaving air temperature. The use of a back pressure valve increases the drop in the suction pressure and it should therefore never be used with hermetic compressors.
(d) Hot gas valve (HGV). The function of this is to stabilise the evaporating pressure (if a self-acting valve is used) or to control the temperature of the air leaving the coil (if a motorised valve is used), by injecting gas from the hot gas line. In the latter case it is possible to achieve tight temperature control if the right quality of valve is chosen. Suction pressure does not fall in either case and hot gas valves are very suitable for use with hermetic compressors. There is only one satisfactory way of using a hot gas valve and that is by injecting the hot gas into the evaporator to impose a false load. Figure 12.8 shows that the evaporator balance can be kept at the point 3 by feeding in enough hot gas to add the false load represented by the vertical distance between the points 3 and P, the suction condition remaining at point 2. The best place to inject the hot gas is into a hot gas header, as Figure 12.8 shows. This then gives the best possible control. An alternative which is cheaper but not so controllable is to put gas into the side of the distributor.
Calculate the mass flow rate of hot gas necessary to control the evaporating pressure at
303.6 kPa for the case considered in Figure 12.8 if the entering wet-bulb is 13°C and a hot gas valve is used instead of a back-pressure valve.
The false load to be imposed on the evaporator is 100 kW — 75 kW = 25 kW. From Table 9.1, liquid Refrigerant 134a at 1°C and 303.6 kPa has a latent heat of 197.9 kJ kg-1 by interpolation. Hence the required mass flow rate is 25/197.9 = 0.126 kg s“1.
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