C L A I M S
1) Refrigerating plant equipped with controlled subcooling comprising at least one compressor (11), an evaporator (12), a thermal expansion valve (13) and a condenser (14) characterised by the insertion of a heat exchanger (15) on the liquid line, upstream of the thermal expansion valve, in which exchanger the liquid is cooled by a lower- temperature external fluid which could be water, air or another refrigerant.
2) Refrigerating plant according to claim 1, characterised in that it uses, as external refrigerant fluid, in addition to air, also tap water or groundwater or water from any other heat reservoir with temperature lower than 22 0 C. 3) Refrigerating plant according to claim 1 or 2, characterised in that said exchanger (15) is of liquid-water type.
4) Refrigerating plant according to claim 1, 2 or 3, characterised in that the calibration valve (16) placed on the water line is commanded by a microprocessor to regulate the power and to optimise the efficiency.
5) Refrigerating plant according to claim 1 , characterised by a two-stage heat recovery, one at low temperature in said exchanger (15) on the liquid line and one at higher temperature in a desuperheater (18) of the refrigerating circuit itself.
6) Refrigerating plant according to claim 5, where said refrigerant fluid which carries out the two-stage heat recovery is water for sanitary use.
7) Refrigerating plant according to claim 1, characterised by a two-stage heat recovery in direct expansion, using a two- stage exchanger (25) on the liquid line.
8) Refrigerating plant according to claim I 5 characterised by the auto-storage of the refrigerating energy, due to the power increase of the subcooling, always tied to the use of the exchanger ( 15) on the line of the liquid.
9) Refrigerating plant according to claim 8, characterised by a functioning in normal conditions in which the pump (Pl) of the hydraulic plant is in function, the pump (P2) of the exchanger (15) is turned off, the valve (V) has path B closed and A open, so that the entire water flow rate passes through the hydraulic store (19) before heading towards the plant.
10) Refrigerating plant according to claim 8, characterised by a functioning in full power conditions in which the store (19) is by-passed, closing the path A of the valve (V) and opening the path B, and the pump (P2) is started and supplies the exchanger (15) with refrigerated water for the controlled subcooling.
11) Refrigerating plant according to claim 8 or 9, characterised in that the cold store is restored shortly before the starting
of the plant.
12) Refrigerating plant according to claim 8 or 9, characterised in that the system works as an ice storage system, without requiring glycol water, and without working at low temperatures with low efficiency.
13) Refrigerating plant according to claim 1, characterised by a controlled subcooling with dedicated refrigerating circuit.
14) Refrigerating plant according to claim 13, characterised in that said exchanger (35) of the subcooling is the evaporator (22) of the second circuit, where said second circuit condenses at the same temperature as the first, but evaporates at a higher temperature.
15) Refrigerating plant according to claim 13 or 14, characterised in that the condenser (24) of the second circuit can also be used for the heat recovery.
Title: Refrigerating plant equipped with controlled subcooling
Applicant: Giuseppe Giovanni Renna
A refrigerating plant equipped with controlled subcooling forms the obj ect of the present finding.
Different embodiments of similar products are known at the state of the art, since it is known that by increasing the subcooling of the liquid at the outlet of the condenser, the performance and energy efficiency (EER) are increased. EER is the ratio between the refrigerating energy provided to the evaporator and the energy absorbed by the compressor. Figure 1 shows the increase of the refrigerant fluid R410A on a pressure - enthalpy diagram (P-H). Similar results can also be obtained with any other refrigerant.
There already exist refrigeration cycles capable of increasing the subcooling with appropriate expedients. There are two currently known cycles: the economizer cycle and the liquid — cold vapour exchanger cycle. The most known economiser circuit is that shown in figure 2, " used both for two-stage centrifugal compressors and for screw compressors equipped with door for the injection of the cold vapour at half compression. At the outlet of the condenser 14 (point 5), the liquid undergoes a first
expansion and is cooled (point 6) to the phase change temperature corresponding to the exit pressure from the first compression stage. Then it enters into the economiser 15 (a heat exchanger), where it transfers heat to the liquid coming from the condenser (point 5), subcooling it up to point 7. The heat taken away from the liquid coming from the condensers ensures that the refrigerant in part evaporates before entering into the second compression stage (in the screw compressors it is in contact with a second suction opening placed at about the midpoint of the rotors). Complete evaporation occurs because the refrigerant is injected in the compressor. During injection, the liquid vaporises, taking away heat: in this manner, the vapour exiting from the first compression stage (point 2) is cooled up to point 3. The economiser on the one hand permits increasing the subcooling, on the other reducing the compression work. It moreover allows limiting the exit temperature of the refrigerant at the end of the compression (point 4): this is very important for the screw compressors which have oil in communication with the high pressure to avoid overly high lubricant temperatures, with consequent cracking phenomena.
A second system for increasing the overheating is given by the refrigerating circuit liquid — cold vapour recovery, shown in figure 3. The circuit is simpler than the preceding. The subcooling increase is obtained by cooling the liquid coming from the condenser 14 by means of the cold vapour coming from
the evaporator 12, which in turn is overheated. Exchanged enthalpy gradient being equal, the thermal gradient on the cold vapour side is about double that on the liquid. The big disadvantage of the already known applications consists of the fact that the subcooling step has very precise thermal limits, which limit its benefits. In the refrigerating circuit with economiser, said limit consists of the flow rate of refrigerant which must be divided between the first and second compression stage and which must reciprocally exchange heat inside the economiser. In the recovery circuit, on the other hand, said limit consists of the need to limit the overheating of the cold vapour coming from the evaporator. In fact, to limit the overheating to 30 0 C, the enthalpy gradient must not be greater than 20 kJ/kg. The obtainable subcooling is therefore only 13 0 C. The technical problem underlying the present invention is therefore that of designing and making a high efficiency refrigerating circuit in which the subcooling step is not subjected to the aforesaid thermal limitations. The finding, object of the present invention, resolves this technical problem since it regards a refrigerating plant equipped with controlled subcooling, characterised by the insertion of a heat exchanger on the liquid line, upstream of the thermal expansion valve. In this exchanger, the liquid is cooled by a lower-temperature external fluid (figure 4) which can be water, air or another refrigerant. The objects and advantages obtained by the invention vary
slightly according to the different adopted embodiments: air + water hybrid cooling; two-stage heat recovery; low-temperature heat recovery by post-heating; auto-storage of refrigerating energy; controlled subcooling with dedicated refrigerating circuit. Said objects and advantages are clarified in the detailed description of the invention which makes specific reference to tables 1/8 - 8/8 in which several preferred, but absolutely non- limiting embodiments are represented. In particular: • Fig. 1 represents the effects of increased subcooling on the performance;
• Fig. 2 shows a refrigerating circuit with economiser of known type;
• Fig. 3 shows a refrigerating circuit with liquid — cold vapour recovery, still of known type;
• Fig. 4 depicts the insertion of the exchanger on the liquid line;
• Fig. 5 is a hybrid condensation system;
• Fig. 6 shows the power required by the plant and output by the refrigerating groups as a function of the air temperature;
• Fig. 7 shows the EER progression for a normal refrigerating group and a hybrid cooling refrigerating group;
• Fig. 8 is the two-stage heat recovery diagram;
• Fig. 9 shows the energy effects of the two-stage heat recovery;
• Fig. 10 shows the output and efficiency improvements in a heat pump due to the two-stage recovery;
• Fig. 11 depicts the energy loss due to the post-heating;
• Fig. 12 shows the application of the controlled subcooling in direct expansion, to recover the post-heating heat;
• Fig. 13 represents the energy advantages of the system as a function of the temperature downstream of the post-heating;
• Fig. 14 shows the auto-storage of the refrigerating energy due to the use of controlled subcooling; • Fig. 15 represents the production of refrigerating energy with the system, object of the present invention;
• Fig. 16 compares the electric power consumptions of the various systems;
• Fig. 17 represents the subcooling controlled by a refrigerating circuit in summer functioning.
With reference to fig. 4, as said, a refrigerating plant is the object of the present finding, equipped with controlled subcooling comprising at least one compressor 11, an evaporator 12, a thermal expansion valve 13 and a condenser 14 and characterised by the introduction, in the refrigeration cycle, of a heat exchanger 15 on the liquid line, upstream of the thermal expansion valve 13. In this exchanger, the refrigerant liquid is cooled by a lower- temperature external fluid. A first embodiment provides for the use, as external refrigerant fluid, of tap water (or groundwater or water from other low-
temperature heat reservoir) which is generally found at 15 0 C, in addition to air. Figure 5 shows the functioning diagram. The exchanger 15 is of liquid- water type. The calibration valve 16 placed on the tap water line is commanded by a microprocessor to regulate the power and to optimise the efficiency. The name of hybrid cooling arises from the fact that air is normally used, but in certain favourable conditions, depending on several factors, in order to increase output and efficiency, diminish costs or reduce noise, the air cooling is supplemented with the water cooling. Figure 6 better specifies the specific and innovative characteristics of the solution. The figure reports the outside air temperature on the x-axis and the refrigerating power required by the plant and that provided by various refrigerating groups on the y-axis. The power required by the plant increases as the outside air temperature increases. Generally, the refrigerating group is chosen to satisfy the power required by the plant in nominal conditions (for example 35 0 C air temperature). In this situation, the normal refrigerating group GF is selected. If, however, the air temperature increases, the plant requires more and the refrigerating group outputs less. There is an imbalance which prevents keeping the plant's thermo-hygrometric conditions under control; it would then be necessary to choose the oversized refrigerating group GF. With the hybrid cooling system, the cooling takes place only with air up to a certain temperature (3O 0 C in the example), after which the water cooling valve is also
opened, and the subcooling is increased, exactly following the plant's requirement curve. The size of the refrigerating group is even less than the normal selection size. The opening of the water cooling can also be commanded based on efficiency, as shown in figure 7. Over a certain air temperature, the valve is opened, improving the efficiency with respect to a normal refrigerating group which is only cooled with air, in greater percentages the higher the outside air temperature. This characteristic leads to a considerable diminution both of consumptions and of the environmental impact of the refrigerating group. Finally, the water cooling can be directed to reduce the noise produced by the refrigerating group. In fact, it is possible to make the fans 17 rotate more slowly, taking advantage of a greater condensation temperature, making up for the diminution of power and efficiency thanks to the hybrid cooling.
A second embodiment provides for a two-stage heat recovery. Generally, the heat recovery occurs at only one temperature level, either by employing all of the condensation heat in a refrigerant - water exchanger parallel to the refrigerant - air exchanger, or by employing only the superheat of the refrigerant in the so-called desuperheater 18. The exchanger 15 can be employed on the liquid line to carry out a two-stage recovery, one with lower temperature in the liquid exchanger and one with higher temperature in the desuperheater. The diagram is shown in figure 8 for the specific case of water for sanitary use, but said
recovery can also be adopted in all cases in which there is a sufficiently low temperature level of the fluid to be heated. This leads to an improvement of the functioning conditions, both in summer (functioning as refrigerator) and in winter (functioning as heat pump).
Let's begin with the summer use. Figure 9 shows the thermal and refrigerating energy requirement of a plant, the electric power consumption as a function of the recovery water requirement in the case of a refrigerating group with total recovery, refrigerating group with partial recovery and two-stage refrigerating group. As seen, the electric power consumption of the two-stage recovery system is always better than the other two in every condition. This is explainable thanks to the clear energy efficiency improvement from the subcooling. For the partial recovery group, the low temperature dashed line of the air indicates that the recovery by the desuperheater was not sufficient, so that it is necessary to carry out an external supplementation (in the figure, supplementing with a heating element is assumed). In winter functioning, as a heat pump, if the liquid - water exchanger is employed to heat the low-temperature water (for example, for the production of hot sanitary water), a double effect is obtained: the increase of the heat pump output and the increase of the winter efficiency, or COP, as shown in figure 10, as a function of the outside air temperature. A third embodiment of the invention, defined heat recovery by
post-heating, is the version of the two-stage heat recovery in direct expansion. In fact, when a cooling is made with post- heating, there is always a loss of energy, as shown in figure 11. The loss is due to the fact that the air is brought to a level of enthalpy lower than that of insertion (point C) to then be heated up to point D. The loss is given by the value δH. There is even loss if the heating from C to D is made by means of a heat recovery by the desuperheater 18. If said recovery is not carried out, the loss substantially doubles (since the necessary heat must be generated). In figure 11, it is seen that the air in point C has a substantially low temperature, so that the subcooling can be increased. Hence, it is possible to recover the post-cooling heat by subcooling. In this manner, the refrigerating output of the system is improved and the loss cancelled. Figure 12 shows how a direct expansion air conditioner can be constructed. If desired, if the post-heating required is high, it is also possible to use the recovery by desuperheater 18, or employ this heat to produce hot sanitary water. The energy savings is quite significant. Figure 13 shows the electric power consumption necessary to bring the air from 28°C with 50% relative humidity up to saturation and then heat it up to the temperature reported on the x-axis. The comparison was made by considering one traditional system with heat recovery by post-heating, and one with electric post-heating. As seen, the system of the invention gives clearly better results. Moreover, the electric post-heating system rapidly increases
consumption due to the energy absorbed by the heating element. The traditional system with hot gas post-heating by desuperheater maintains constant consumption while there is the possibility to employ the desuperheater heat, after which also in this case a heating element must be used. The invention, instead, decreases the required power. In fact, the greater the required post-heating load, the greater the output of the refrigerating group, consumption being equal, since it increases the subcooling. Consequently, consumptions decrease. A fourth embodiment of the invention, certainly the most interesting application, is the auto-storage of refrigerating energy. The power increase of the subcooling is utilised to increase the possible refrigerating energy in the hydraulic store 19 (fig.14). There is an over 10-times increase of the stored energy, because one can employ the over 3O 0 C thermal gradient as opposed to a 3 0 C thermal gradient (from 7 to 10 0 C). Figure 14 better illustrates the functioning. In normal functioning, pump Pl of the hydraulic plant is in function, pump P2 of the exchanger is turned off, the valve V has path B closed and path A open. The entire water flow rate passes through the hydraulic store 19 before heading towards the plant. When the refrigerating group works at full power, the store is by-passed, closing path A of the valve V and opening path B. At this point, the pump P2 starts, which feeds the exchanger 15 with water cooled to 7°C for the controlled subcooling. The obtainable increase is greater than
40%, electrical consumption being equal. When the plant is turned off, the temperature of the water store can be restored. Therefore, there is an energy production at different times from that of maximum consumption. The water store substantially works as an energy store. Since the thermal gradient obtainable on the subcooling exchanger is about 35°C (7°C - 42°C), the storable energy is about 40 kWh for every 1000 litres of water. If the same water content was used directly for the plant, one could utilise a thermal gradient of only 3.5 0 C, only storing 4 kWh every 1000 litres. In such a manner, by appropriately sizing the water stores and adapting them to the power of the refrigerating group, a thermal store system is obtained. In this manner, the electric energy requirement and the size of the refrigerating group can be decreased. As an example, 145 kW of maximum power can be obtained with a refrigerating group from only 100 kW with an electric power absorption of only 30 kW. Figure 15 shows the progression of the power production with respect to the load requirement. As seen, the refrigerating group of the present invention gives at most 100% of instantaneous power, while the remaining part is obtained from the increase of the subcooling through the store. The cold store is restored shortly before the starting of the plant. As seen, the system works as an ice storage system, with the unquestionable advantage of requiring neither glycol water nor working at low temperatures with low efficiency. In fact, as seen in figure 16, with respect to
an ice storage plant, the consumptions are reduced during the preparation of the store. The hydraulic circuit also results considerably simplified.
The fifth and final embodiment of the invention provides for a controlled subcooling with dedicated refrigerating circuit. It is the most complex and least applicable system and can give rather good results in a heat pump. The additional subcooling is obtained through a second refrigerating circuit. The exchanger of the subcooling 35 (fig. 17) is on one side the evaporator 22 of the second circuit. The second circuit condenses at the same temperature of the first, but evaporates to a higher temperature. In summer, there are two advantages: on one hand the second compressor 21 is smaller with respect to the increase of power provided to the first compressor, since it evaporates at higher temperature. On the other hand, the overall efficiency is increased because the second circuit works with an improved EER. Figure 17 shows the system in summer functioning. The condenser 24 of the second circuit can also be used for the heat recovery. This appears very interesting if a CO 2 refrigeration cycle is used for the production of high temperature sanitary water. In this case, the production would also increase the refrigerating output and efficiency. The heat pump functioning seems more interesting. Both condensers work to produce hot water and condense at the same temperature, but the evaporation in the second circuit occurs at a much higher temperature, which
clearly improves the overall COP, particularly with lower outside air temperatures.
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