DESCRIPTION

BACKGROUND OF THE INVENTION

Field of the Invention

The present invention is applicable in the energy industry, for the conversion of heat sources at low-, medium- or high-temperature, which allows generating energy in the turbine with high energy efficiency, said energy being mechanical or electrical energy, in the latter case when the turbine is coupled to an electric generator.

Description of the Related Art

Supercritical carbon dioxide regenerative Brayton cycle with multiple recuperators and auxiliary compressors (referred to as "multiple recompression cycle" from now on) improves the energy efficiency of the conversion of thermal energy from low-, medium- and high-temperature heat sources to mechanical or electrical energy when compared to state-of-the-art regenerative Brayton recompression cycle.

In some methods used today (such as state-of-the-art recompression cycles as the one described in D. Novales, A. Erkoreka, V. De la Pena, B. Herrazti, Sensitivity analysis of supercritical CO2 power cycle energy and exergy efficiencies regarding cycle component efficiencies for concentrating solar power, Energy Conversion and Management 182 (2019) 430-450.), a maximum number of two recuperators are used in series to recover the heat from the turbine outlet. However, the heat recovery of this cycle is not optimized.

Summary of the invention

The object of the present invention is to improve the energy efficiency of supercritical carbon dioxide recompression cycles of the state-of-the-art through the use of a new cycle configuration that improves the heat recovery process.

Any heat source could be used for the cycle, such as heat of solar origin or nuclear origin, heat obtained from the combustion of matter such as fossil fuels, biomasses, waste or biogas, waste heat coming from any process or any other heat source which reaches the temperatures required in the present invention.

For cycle definition, the first step is to define the optimum number of recuperators N to be installed in the cycle (where N > 3). The optimum number of recuperators in the cycle is calculated as follows.

First, Turbine Inlet (stream Tl according to Figure 1) pressure (P _TI ) and temperature (T _TI ) and turbine outlet (stream RHIN according to Figure 1) pressure (P _RHIN ) are defined. RHI _N stands for Recuperator N Hot Inlet, see Figure 1. Also, turbine isentropic efficiency (η _{s, T} ) is defined. Then, turbine outlet temperature is obtained using thermophysical properties of CO ₂ and turbine isentropic efficiency definition as per equations (1) to (5). Where, h _TI stands for the specific enthalpy of stream Tl, s _TI stands for the specific entropy of stream stands for the specific enthalpy of stream RHIN for an adiabatic and isentropic expansion, stands for the specific enthalpy of stream RHIN.

Also, main compressor inlet (stream MCI according to Figure 1) pressure (P _in ) and temperature (T _in,1 ) are defined. Main compressor inlet pressure shall be the same or lower than turbine outlet pressure (pressure at state RHIN according to Figure 1). Main compressor outlet pressure (P _out ) shall be the same or higher than turbine inlet pressure (pressure at state Tl according to Figure 1). Isentropic efficiency of each compressor is defined as η _s,C,i , being i the number of the compressor, numbering the compressors from the left to the right in Figure 1 (i.e., for main compressor i=1 , for auxiliary compressor 1 i=2, etc.). Then, equations (6) to (10) are applied to determine the outlet temperature of the i ^th compressor (T _out,i ), where i is defined as the number of times equations (6) to (10) have been used (the iteration number), starting with i = 1 for the first iteration. T _{in i} stands for the inlet temperature of the i ^th compressor, h _{in i} stands for the inlet specific enthalpy of the i ^th compressor, s _inii stands for the inlet specific entropy of the i ^th compressor, h _{out s i} stands for the outlet specific enthalpy of the of the i ^th compressor for an adiabatic and isentropic compression, h _out,i stands for the specific enthalpy of the stream leaving the i ^th compressor, T _in,1 is defined as the temperature of stream MCI and pinch _i is defined as the minimum temperature difference between the cold stream and hot stream of the i ^th recuperator.

Once we get T _out,i , if equation (11) is not fulfilled, equations (6) to (10) are again applied using T _in,i = T _out,i-1 + pinch _i-1 and P _in as starting point of the new iteration.

The optimum number of recuperators N that can be included in the present invention cycle is defined as i-1 , being i the number of the iteration when the stopping criterion (equation 11) is complied with. One main compressor and N-1 auxiliary compressors are associated with those N recuperators according to the configuration shown in Figure 1.

In the event that the stopping criterion (equation 11) occurs when T _{out i} « T _RHIN - pinchi, ^a particular case is obtained in which N recuperators, one main compressor and N auxiliary compressors constitute the optimal configuration.

For example, for a turbine inlet pressure of 20 MPa, a turbine inlet temperature of

680 °C, a main compressor inlet pressure of 7.5 MPa, an isentropic efficiency of the turbine of 93% and an isentropic efficiency of all compressors of 95%, the optimum number of recuperators to be included in the cycle is 4.

Obviously, this procedure calculates the optimum number of recuperators with their corresponding auxiliary compressors that could be installed in the cycle to maximize the cycle efficiency. However, a higher or lower number of recuperators with their auxiliary compressors can also be installed if required, but cycle efficiency will decrease for some given turbine inlet temperature and pressure, and turbine outlet pressure.

According to the invention, once defined the optimum number of recuperators of the cycle, N, as indicated previously, the method for generating energy by means of a multiple recompression cycle using supercritical carbon dioxide (sCO ₂ ) as a working fluid comprises the following steps, according to the numbering indicated in Figure 1 :

1) sCO ₂ stream at the Recuperator N Cold Outlet ( RCO _N ) is at pressures between 7.5 MPa and 50 MPa and is heated by means of an external heat source to temperatures between 50 °C and 900 °C, reaching stream Turbine Inlet (Tl).

2) sCO ₂ stream Tl is expanded in a turbine to a pressure between 3 MPa and 10 MPa (stream Recuperator N Hot Inlet (RH IN)), and generates some mechanical or electrical energy, in the latter case when the turbine is coupled to an electric generator.

3) sCO ₂ stream RH IN is cooled in N recuperators. The optimal number of total recuperators N is calculated as described above. The total number of recuperators N are put in series, one after the other.

4) sCO ₂ stream RH IN is cooled to stream RHON (Recuperator N Hot Outlet) and heats stream RCI _N (Recuperator N Cold Inlet) to RCO _N in the recuperator number N.

5) sCO ₂ stream RH I _N-1 is cooled to stream RHO _N-1 in recuperator number N-1 and heats stream RCI _N-1 to RCO _N-1 . Stream RHO _N-1 is split into two streams: RH I _N-2 and Auxiliary Compressor N-1 Inlet (ACI _N-1 ). Stream ACI _N-1 is compressed in auxiliary compressor N-1 to, preferably, the pressure of stream RCO _N-1 generating the stream ACO _N-1 (Auxiliary Compressor N-1 Outlet). Stream ACO _N-1 is mixed with stream RCO _N-1 , obtaining from the mixture stream RCI _N . Stream RH IN-2 is sent to recuperator N-2.

6) If N>3, the pattern shown in step 5) is repeated for recuperators N-2 to 2.

7) For recuperator 1 , the sCO ₂ stream RHI ₁ is cooled to stream RHO1 in recuperator number 1 and heats stream RCI ₁ to RCO ₁ . Stream RHO1 is split into two streams: Cl (Cooler Inlet) and Auxiliary Compressor 1 Inlet (ACI ₁ ). Stream ACI ₁ is compressed in auxiliary compressor 1 to, preferably, the pressure of stream RCO ₁ generating the stream ACO1. Stream ACO1 is mixed with stream RCO ₁ , obtaining from the mixture stream RCI2. Stream Cl is sent to the Cooler.

8) Preferably, to make the heat recovery optimal, the split factors (Oj) must ensure a similar heat rate capacity in both streams of each recuperator and a similar pinch value in the hot and cold section of each recuperator. The split factor (Oj) is defined as the ratio of ACIj mass flow rate divided by the total CO2 mass flow being expanded in the turbine (stream Tl).

9) Preferably, the temperature of each ACOj stream must be similar to the temperature of the corresponding RCOj stream with which is mixed.

10) Preferably, the pinch values of all recuperators must be similar between them.

11) sCO ₂ stream Cl (Cooler Inlet) is cooled down to a temperature between -10 °C and 70 °C (preferably 32 °C) using any external cooling sink, reaching stream MCI (Main Compressor Inlet).

12) sCO ₂ stream MCI is compressed in the main compressor to the same or higher pressure than the one defined for the turbine inlet (stream Tl). The outlet of the main compressor is stream RCI ₁ .

13) Stream RCI ₁ is heated to stream RCO ₁ in recuperator 1 using heat transferred from stream RHI ₁ to stream RHO ₁ . Then, the stream RCO ₁ is mixed with stream ACO1 and the mixture is stream RCI ₂ .

14) Pattern shown in step 13) is repeated for recuperators 2 to N-1.

15) Stream RCI _N is heated to stream RCO _N in recuperator N by cooling stream RHIN to stream RHO _N .

The multiple recompression cycle of the invention improves the efficiency of the state- of-the-art recompression cycle, which has only two recuperators and one auxiliary compressor. An example of this increase in efficiency can be seen in Table 1. The comparison of these two cycles allows a better understanding of the key aspect of the present invention with respect to the heat recovery in the recuperators. Figure 2 presents the schematic diagram of the state-of-the-art recompression cycle. Figure 3 represents the Temperature-Thermal Power Exchange diagram within the two recuperators of the Figure 2 recompression cycle for the Table 1 example. The state-of-the-art recompression cycle configuration does not permit a proper heat recovery of the sCO ₂ stream leaving the turbine at 548 °C. It only permits heating the high pressure sCO ₂ stream up to 506 °C (stream 14 of Figure 2). The separation of the temperature profiles occurs in both recuperators (see Figure 3), but it is more notorious in the high temperature recuperator.

Following the above presented steps 1) to 15), for 680 °C and 20 MPa inlet conditions on the turbine and 7.5 MPa in the outlet, the configuration of the present invention presented in Figure 4 is obtained. Table 1 compares cycles with equal isentropic efficiencies in turbomachinery and equal effectiveness values in all the heat exchangers. Four recuperators and three auxiliary compressors are the optimal configuration following the aforementioned steps. As can be seen in Figure 5 Temperature-Thermal Power Exchange diagram, for the same outlet turbine temperature as for the recompression cycle (548 °C), the heat recovery on the present invention permits to reach 537 °C (stream 14 of Figure 4) in the high pressure sCO ₂ stream.

Even if the compression work is higher in the present invention case, in Table 1 it can be seen that efficiency improves by 3.84 points, thanks to a more efficient heat recovery in the recuperators due to the configuration of the present invention. This improvement in efficiency allows increasing the electrical or mechanical energy generated per operating hour for the same heat input.

Table 1- Comparison between recompression cycle and present invention

The possibility of using at least one intercooling stage in the main compression process is contemplated. In the cycles of the state-of-the-art such as the recompression cycle, the use of one or various intercooling stages in the main compression process reduces the compression work, but it makes the heat recovery process more irreversible, mainly in the high temperature recuperator. This is because the high temperature recuperator needs to exchange more heat when the intercooling is present. Since the heat capacity rate of the hot side stream in the high temperature recuperator is lower than the heat capacity rate of the cold side stream, the additional heat required to be exchanged in the high temperature recuperator, generates a higher temperature difference between the streams in the hot section of this recuperator. Consequently, more irreversibilities are present in the high temperature recuperator and this effect reduces the efficiency of the cycle. The effect of the efficiency increase due to compression work reduction does not always compensate the efficiency reduction due to a worse heat recovery (as shown in Table

2).

Table 2- Comparison between recompression cycle and present invention with one intercooling stage

Table 2 presents the same cycles as Table 1 but including an interceding stage in the main compression process. In Table 2 it can be seen that the recompression cycle efficiency is reduced from 53.42% (Table 1) to 52.64% (Table 2) due to the inclusion of the intercooling stage. On the other hand, introducing the same intercooling stage to the present invention cycle permits to increase the thermal efficiency of the cycle from 57.26% (Table 1) to 58.05% (Table 2). The intercooling stage on the present invention can be seen in Figure 6.

As proven above, the proposed strategy of using N recuperators in the cycles of the present invention reduces the inefficiencies of the heat recovery process due to the use of intercooling in the main compression process. The use of at least one intercooling stage in the main compression process reduces the temperature of RCI ₁ stream (see Figure 1 and stream 1 of Figure 6) and consequently the RHOi temperature (Figure 1) is also reduced, increasing the total heat recovery while maintaining a similar RCO _N stream temperature (Figure 1). This is obtained by means of the following steps:

16) Instead of using one main compressor to compress the MCI stream directly to RCI ₁ stream, the main compression process is performed with at least one interceding stage and then steps 1) to 15) are applied to define the multiple recompression cycle.

When intercooling is used, the application of equations (6) to (11) has to consider the intercooling effect on the main compression process for the calculation of T _out,1 (this is the stream RCI ₁ of Figure 1). Since T _out,1 with interceding is lower than T _out,1 without intercooling, the iterative process of equations (6) to (11) to determine the optimum number of recuperators, may lead to a multiple recompression cycle with more than N recuperators than the cycle without intercooling. Once the optimal number of N recuperators is calculated for the cycle with intercooling using the method explained previously for some specific CO2 conditions and equipment specifications, steps 1) to 15) are applied to define the multiple recompression cycle. As shown in Table 3, for medium temperature heat sources the efficiency of the multiple recompression cycle is higher than the efficiency of the state-of-the-art cycles, with an increase of efficiency of 0.94 points. Figure 8 presents the schematic of the multiple recompression cycle obtained from the fulfilment of the steps above for the medium temperature turbine inlet conditions presented in Table 3. Figure 9, presents the Temperature-Thermal Power Exchange diagram for the present invention case of Table 3.

Table 3- Comparison between state-of-the-art water-steam cycle and present invention cycle for a medium temperature solar power plant

As presented in Table 4, for low temperature heat sources, the efficiency of the multiple recompression cycle is higher than the efficiency of the state-of-the-art Organic Rankine Cycles (ORC cycles), with an increase of efficiency of 2.1 points. Figure 10 shows the schematic of the multiple recompression cycle obtained from the fulfilment of the steps above for the turbine inlet and outlet conditions presented in Table 4. Figure 11 presents the Temperature-Thermal Power Exchange diagram for the present invention case of Table 4.

Table 4- Comparison between state-of-the-art ORC cycle and present invention cycle for a low temperature Internal Combustion Engine cooling water heat recovery

As shown in Table 5, for high-temperature heat sources with the availability of a cold sink that allows the CO2 to cool down to temperatures below its critical temperature, the efficiency of the multiple recompression cycle is 1.23 points greater than the efficiency of the recompression cycle. In this case, the CO2 is expanded to a subcritical pressure of 5.3 MPa and the turbine inlet pressure is increased to 35 MPa to take advantage of heat sources in the form of hot mass flows that require to cool down about 240 °C. In this way, the pressure jump available in the turbine allows the CO2 to be cooled through an expansion from 680 °C to 437 °C. Figure 12 presents the diagram of the optimal multiple recompression cycle obtained from the fulfilment of the steps described above for the turbine inlet and outlet conditions presented in Table 5. Figure 13 presents the Temperature-Thermal Power Exchange diagram during the heat recovery process for the present invention case in Table 5. Table 5- Comparison between the recompression cycle and the present invention fora cold sink that allows the CO2 to be cooled to temperatures below the critical temperature and a hot source in the form of a hot mass flow that requires a thermal jump of about 240 °C.

If a heat transfer fluid or a hot stream is used as the heat source, the use of one or various steps of reheating is also contemplated for the expansion process of the present invention cycle. However, if the temperature profiles on both sides of the heat source heat exchanger are parallel to each other, meaning that the heat rate capacity of the fluids in both sides of the heat source heat exchanger are similar, the inclusion of the reheating has a negligible effect on the increase of the energy efficiency of the cycle.

Description of the Drawings

To complement the description which is being made and for the purpose of aiding to better understand the features of the invention according to a preferred practical embodiment thereof, a set of drawings is attached as an integral part of said description, in which the following has been depicted with an illustrative and nonlimiting manner:

Figure 1- Schematic diagram of the multiple recompression cycle with N recuperators. Figure 2- Schematic diagram of the state-of-the-art recompression cycle.

Figure 3- Temperature - Thermal Power Exchange diagram of the heat recovery process within the two recuperators of the state-of-the-art recompression cycle for a high temperature heat source.

Figure 4- EMBODIMENT 1 , schematic diagram of the multiple recompression cycle with four recuperators.

Figure 5- Temperature - Thermal Power Exchange diagram of the heat recovery process within the four recuperators of the multiple recompression cycle for a high temperature heat source. Figure 6- EMBODIMENT 2, schematic diagram of the multiple recompression cycle with four recuperators and one intercooling stage.

Figure 7- Temperature - Thermal Power Exchange diagram of the heat recovery process within the four recuperators of the multiple recompression cycle including one intercooling stage for a high temperature heat source.

Figure 8- EMBODIMENT 3, schematic diagram of the multiple recompression cycle with three recuperators.

Figure 9- Temperature - Thermal Power Exchange diagram of the heat recovery process within the three recuperators of the multiple recompression cycle for a medium-temperature heat source.

Figure 10- EMBODIMENT 4, schematic diagram of the multiple recompression cycle with three recuperators and three auxiliary compressors.

Figure 11- Temperature - Thermal Power Exchange diagram of the heat recovery process within the three recuperators of the multiple recompression cycle for a low- temperature heat source.

Figure 12- EMBODIMENT 5, schematic diagram of the multiple recompression cycle with three recuperators.

Figure 13- Temperature - Thermal Power Exchange diagram of the heat recovery process within the three recuperators of the multiple recompression cycle for a high- temperature heat source and a cold sink that allows the CO2 to be cooled to temperatures below its critical temperature.

Preferred embodiment of the invention

As has been set forth, the invention comprises combinations of several elements which have synergistic effects on the improvement of the energy efficiency and on the use of different heat source temperature ranges. Five embodiments are described below, without these examples being a limitation to the possibilities of combination and application of the inventive concepts described above.

Figure 4 shows a highly regenerative Brayton cycle with multiple recuperators and auxiliary compressors driven by a high-temperature heat source stream.

The cycle depicted in said Figure 4 is a preferred embodiment of the invention for electric generation by means of a heat source available at high temperature. This preferred embodiment, has four recuperators and three auxiliary compressors. It must be noted that, from now on, when making reference to the total sCO ₂ mass flow rate, total sCO ₂ mass flow being expanded in the turbine is being referred.

In view of said Figure 4, the high temperature heat source permits to heat up the sCO ₂ stream leaving the recuperator 4 (stream 14) up to 680 °C at 20 MPa (stream 15). The stream 15 is expanded in the turbine to 548 °C and about 7.5 MPa (stream 16).

Stream 16 enters the hot side of recuperator 4 and is cooled down to 428.5 °C (stream 19) by means of heating stream 10 from 422 °C to 537 °C (stream 14). Stream

19 is then cooled down in the recuperator 3 to 308 °C (stream 20) by heating stream 7 from 301 .5 °C to 421 .5 °C (stream 8). Auxiliary compressor 3 compresses the 6.4% of the total sCO ₂ mass flow rate from about 7.5 MPa and 308 °C to about 20 MPa and 429 °C (stream 9). Stream 9 is mixed with stream 8 to obtain the stream 10 mentioned above. The 93.6% of total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 2 at about 7.5 MPa and 308 °C (stream 21).

Stream 21 is then cooled down in the recuperator 2 to 191 .5 °C (stream 22) by heating stream 4 from 185.5 °C to 302 °C (stream 5). Auxiliary compressor 2 compresses the 13.2% of the total sCO ₂ mass flow rate from about 7.5 MPa and 191.5 °C to about

20 MPa and 299 °C (stream 6). Stream 6 is mixed with stream 5 to obtain stream 7. The 80.4% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 1 at about 7.5 MPa and 191.5 °C (stream 23).

Stream 23 is then cooled down in the recuperator 1 to 90 °C (stream 24) by heating stream 1 from 85 °C to 187 °C (stream 2). Auxiliary compressor 1 compresses the 28.8% of the total sCO ₂ mass flow rate from about 7.5 MPa and 90 °C to about 20 MPa and 183 °C (stream 3). Stream 3 is mixed with stream 2 to obtain stream 4. The 51 .6% of the total sCO ₂ mass flow rate goes to the cooler at about 7.5 MPa and 90 °C (stream 25).

Stream 25 is cooled in the cooler from about 90 °C to about 32 °C (stream 26). Stream 26 is compressed in the main compressor from about 32 °C and 7.5 MPa to about 85 °C and 20 MPa (stream 1).

This embodiment allows achieving increases up to 3.8 points with respect to the state- of-the-art recompression cycle working with equipment with identical isentropic efficiencies and effectiveness. Said Figure 4 shows a preferred embodiment for the exploitation of a heat source at high temperature.

On the other hand, according to a second embodiment, Figure 6 shows a multiple recompression cycle that uses a high temperature heat source with an intermediate cooling stage in the main compression process.

The cycle depicted in said Figure 6 is a preferred embodiment of the invention for electric generation by means of a heat source available at high temperature. This preferred embodiment, has four recuperators, three auxiliary compressors and one intercooling stage in the main compression process.

In view of said Figure 6, the high temperature heat source permits to heat up the sCO ₂ stream leaving the recuperator 4 (stream 14) up to 680 °C at 20 MPa (stream 15). The stream 15 is expanded in the turbine to 548 °C and about 7.5 MPa (stream 16). Stream 16 enters the hot side of recuperator 4 and is cooled down to 398 °C (stream 19) by means of heating stream 10 from 390 °C to 534 °C (stream 14). Stream 19 is then cooled down in the recuperator 3 to 264 °C (stream 20) by heating stream 7 from 257 °C to 390.5 °C (stream 8). Auxiliary compressor 3 compresses the 8.2% of the total sCO ₂ mass flow rate from about 7.5 MPa and 264 °C to about 20 MPa and 380 °C (stream 9). Stream 9 is mixed with stream 8 to obtain stream 10. The 91 .8% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 2 at about 7.5 MPa and 264 °C (stream 21).

Stream 21 is then cooled down in the recuperator 2 to 150 °C (stream 22) by heating stream 4 from 144 °C to 258 °C (stream 5). Auxiliary compressor 2 compresses the 18.4% of the total sCO ₂ mass flow rate from about 7.5 MPa and 150 °C to about 20 MPa and 252 °C (stream 6). Stream 6 is mixed with stream 5 to obtain stream 7. The 73.4% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 1 at about 7.5 MPa and 150 °C (stream 23).

Stream 23 is then cooled down in the recuperator 1 to 56 °C (stream 24) by heating stream 1 from 52 °C to 146 °C (stream 2). Auxiliary compressor 1 compresses the 29.3% of the total sCO ₂ mass flow rate from about 7.5 MPa and 56 °C to about 20 MPa and 140 °C (stream 3). Stream 3 is mixed with stream 2 to obtain stream 4. The 44.1% of the total sCO ₂ mass flow rate goes to the cooler about 7.5 MPa and 56 °C (stream 25).

Stream 25 is cooled in the cooler from about 56 °C to about 32 °C (stream 26). Stream 26 is compressed in the main compressor 1 from about 32 °C and 7.5 MPa to about 59 °C and 12.25 MPa (stream 27). Stream 27 is cooled to about 40 °C in the intercooler to obtain stream 28. Stream 28 is compressed in main compressor 2 to about 20 MPa and 52 °C (Stream 1). This embodiment allows achieving increases up to 4.6 points with respect to the state- of-the-art recompression cycle without intercooling working with equipment with identical isentropic efficiencies and effectiveness. Note that this embodiment allows achieving increases up to 5.4 points with respect to the state-of-the-art recompression cycle with intercooling working with equipment with identical efficiencies and effectiveness and an identical intercooling stage. Figure 6 shows a preferred embodiment for the exploitation of a heat source at high temperature.

Likewise, according to a third preferred embodiment, depicted in Figure 8 there is a multiple recompression cycle using three recuperators and two auxiliary compressors. The cycle depicted in said Figure 8 is a preferred embodiment of the invention for electric generation by means of a heat source available at medium temperature.

In view of said Figure 8, the medium temperature heat source permits to heat up the sCO ₂ stream leaving the recuperator 3 (stream 14) up to 377 °C at 17 MPa (stream

15). The stream 15 is expanded in the turbine to 289 °C and about 7.5 MPa (stream

16).

Stream 16 enters the hot side of recuperator 3 and is cooled down to 240 °C (stream 20) by means of heating stream 7 from 238 °C to 282 °C (stream 14). Stream 20 is then cooled down in the recuperator 2 to 160 °C (stream 22) by heating stream 4 from 156 °C to 236 °C (stream 5). Auxiliary compressor 2 compresses the 16.3% of the total sCO ₂ mass flow rate from about 7.5 MPa and 160 °C to about 17 MPa and 246 °C (stream 6). Stream 6 is mixed with stream 5 to obtain stream 7. The 83.7% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 1 at about 7.5 MPa and 160 °C (stream 23).

Stream 23 is then cooled down in the recuperator 1 to 80 °C (stream 24) by heating stream 1 from 76 °C to 156.5 °C (stream 2). Auxiliary compressor 1 compresses the 32.5% of the total sCO ₂ mass flow rate from about 7.5 MPa and 80 °C to about 17 MPa and 155.5 °C (stream 3). Stream 3 is mixed with stream 2 to obtain stream 4. The 51.2% of the total sCO ₂ mass flow rate goes to the cooler at about 7.5 MPa and 80 °C (stream 25).

Stream 25 is cooled in the cooler from about 80 °C to about 32 °C (stream 26). Stream 26 is compressed in the main compressor from about 32 °C and 7.5 MPa to about 76 °C and 17 MPa (stream 1).

This embodiment allows achieving increases up to 0.94 points with respect to the state-of-the-art water-steam regenerative Rankine cycle. Said Figure 8 shows a preferred embodiment for the exploitation of a heat source at medium temperature. In this case, the cold outlet temperature of the heat source stream is fixed by the solar field. The selected turbine inlet pressure permits to work with the Heat Transfer Fluid entering the Heat Transfer Fluid Heat Exchanger at about 390 °C (stream HSi) and leaving this exchanger at about 295 °C (stream HS2).

Besides, according to a fourth preferred embodiment depicted in Figure 10, there is a multiple recompression cycle using three recuperators and three auxiliary compressors. The cycle depicted in said Figure 10 is a preferred embodiment of the invention for electric generation by means of a heat source available at low temperature.

In view of said Figure 10, the low temperature heat source permits to heat up the sCO ₂ stream leaving the recuperator 3 (stream 14) up to 85 °C at 8.6 MPa (stream

15). The stream 15 is expanded in the turbine to 73.7 °C and about 7.5 MPa (stream

16).

Stream 16 enters the hot side of recuperator 3 and is cooled down to 61.9 °C (stream 20) by means of heating stream 7 from 61.35 °C to 73.05 °C (stream 8). Auxiliary compressor 3 compresses the 15% of the total sCO ₂ mass flow rate from about 7.5 MPa and 61.9 °C to about 8.6 MPa and 73.4 °C (stream 9). Stream 9 is mixed with stream 8 to obtain the stream 14 at 73.1 °C and 8.6 MPa. The 85% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 2 at about 7.5 MPa and 61.9 °C (stream 21).

Stream 21 is then cooled down in the recuperator 2 to 50.4 °C (stream 22) by heating stream 4 from 49.9 °C to 61.3 °C (stream 5). Auxiliary compressor 2 compresses the 20.4% of the total sCO ₂ mass flow rate from about 7.5 MPa and 50.4 °C to about 8.6 MPa and 61.5 °C (stream 6). Stream 6 is mixed with stream 5 to obtain stream 7. The 64.6% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 1 at about 7.5 MPa and 50.4 °C (stream 23).

Stream 23 is then cooled down in the recuperator 1 to 39.8 °C (stream 24) by heating stream 1 from 39.4 °C to 49.8 °C (stream 2). Auxiliary compressor 1 compresses the 33.6% of the total sCO ₂ mass flow rate from about 7.5 MPa and 39.8 °C to about 8.6 MPa and 50.0 °C (stream 3). Stream 3 is mixed with stream 2 to obtain the stream 4. The 31.0% of the total sCO ₂ mass flow rate goes to the cooler at about 7.5 MPa and 39.8 °C (stream 25).

Stream 25 is cooled in the cooler from about 39.8 °C to about 32 °C (stream 26). Stream 26 is compressed in the main compressor from about 32 °C and 7.5 MPa to about 39.4 °C and 8.6 MPa (stream 1).

This embodiment allows achieving increases up to 2.1 points with respect to the state- of-the-art ORC cycles. Said Figure 10 shows a preferred embodiment for the exploitation of a heat source at low temperature being the cold outlet temperature of the heat source stream (stream HS2) fixed by the heat source stream cooling requirements. In this embodiment, the selection of 8.6 MPa as the turbine inlet pressure leads to a particular case where there are as many recuperators as auxiliary compressors.

Finally, according to a fifth preferred embodiment, depicted in Figure 12 there is a multiple recompression cycle using three recuperators and two auxiliary compressors. The cycle depicted in said Figure 12 is a preferred embodiment of the invention for electric generation by means of a high-temperature heat source and a cold sink that allows the CO2 to be cooled to temperatures below its critical temperature. This configuration makes it possible to take advantage of hot sources in the form of mass flows or hot streams that must be cooled about 240 °C by expanding the sCO2 from 35 MPa to subcritical pressures of 5.3 MPa.

In view of said Figure 12, the high temperature heat source permits to heat up the sCO ₂ stream leaving the recuperator 3 (stream 14) up to 680 °C at 35 MPa (stream

15). The stream 15 is expanded in the turbine to 437 °C and about 5.3 MPa (stream

16).

Stream 16 enters the hot side of recuperator 3 and is cooled down to 391 °C (stream 20) by means of heating stream 7 from 389 °C to 430 °C (stream 14). Stream 20 is then cooled down in the recuperator 2 to 200 °C (stream 22) by heating stream 4 from 189 °C to 381 °C (stream 5). Auxiliary compressor 2 compresses the 20.3% of the total sCO ₂ mass flow rate from about 5.3 MPa and 200 °C to about 35 MPa and 420 °C (stream 6). Stream 6 is mixed with stream 5 to obtain stream 7. The 79.7% of the total sCO ₂ mass flow rate goes to the hot side inlet of recuperator 1 at about 5.3 MPa and 200 °C (stream 23).

Stream 23 is then cooled down in the recuperator 1 to 32 °C (stream 24) by heating stream 1 from 26.5 °C to 186 °C (stream 2). Auxiliary compressor 1 compresses the 25.6% of the total sCO ₂ mass flow rate from about 5.3 MPa and 32 °C to about 35 MPa and 197.5 °C (stream 3). Stream 3 is mixed with stream 2 to obtain stream 4. The 54.1 % of the total sCO ₂ mass flow rate goes to the cooler at about 5.3 MPa and 32 °C (stream 25).

Stream 25 is cooled in the cooler from about 32 °C to about 5 °C (stream 26). Stream 26 is compressed in the main compressor from about 5 °C and 5.3 MPa to about 26.5 °C and 35 MPa (stream 1).

This embodiment allows achieving increases up to 1 .2 points with respect to the state- of-the-art recompression cycle working with equipment with identical isentropic efficiencies and effectiveness. Said Figure 12 shows a preferred embodiment for the exploitation of a heat source at high temperature and a cold sink that allows the CO2 to be cooled to temperatures below its critical temperature. In this case, the outlet temperature of the hot stream or thermal fluid that works as a heat source would be set at about 460 °C. The selected turbine inlet pressure permits to work with the Heat Transfer Fluid entering the Heat Transfer Fluid Heat Exchanger at about 700 °C (stream HS1) and leaving this exchanger at about 460 °C (stream HS2).