ES2657072A1 - THERMODYNAMIC CYCLIC PROCESS WITH TURBINE AND COMPRESSOR OF GAS, WITH CONTRIBUTION OF HEAT BY EXTERNAL SOURCE, AND DEVICE FOR ITS REALIZATION (Machine-translation by Google Translate, not legally binding) - Google Patents

Abstract

Thermodynamic cyclic process with turbine and gas compressor, with heat supply by external source, and device for its realization, consisting of a working fluid cycle that is continuously maintained in a gaseous state, and receives heat from an external source, just before entering the turbine, from which it emerges to enter a regenerative exchanger balanced in specific heat, with little cold acting before the compressor, whose output constitutes the high-pressure circuit of the regenerator. The fluid can be a pure substance, or a mixture to improve the joint thermodynamic properties. (Machine-translation by Google Translate, not legally binding)

Description

image 1

THERMODYNAMIC CYCLIC PROCESS WITH TURBINE AND GAS COMPRESSOR, WITH SUPPLY OF HEAT FROM AN EXTERNAL SOURCE, AND DEVICE FOR ITS REALIZATION

DESCRIPTION

TECHNICAL SECTOR

The invention falls within the field of thermodynamic cycles that transform thermal energy into kinetic energy of the axis of rotation of its expanding machine or turbine.

TECHNICAL PROBLEM TO BE SOLVED AND BACKGROUND OF THE INVENTION

The problem consists of making the most of the energy of an external heat source, devising a thermodynamic assembly that uses relatively conventional machines and equipment, but interconnected in a novel way, exploiting the thermo-physical qualities of the real working fluid, obtaining results that they go beyond the state of the art.

From the theoretical point of view and analysis of proposals, the state of the art can be seen described in the previous applications of the inventor of the present application; Specifically, patent ES 2427648 B2 deals with a Brayton cycle with ambient cooling close to the critical isotherm, the second document being, with application number of the national patent application P201731263, a cycle with a point of lower enthalpy that has a temperature of below criticism, but his pressure is above critical pressure.

The technological and commercial exploitation of gas turbines has focused for many years on jet turboprops, although at the end of the 20th century the gas turbine began to prevail as the head of a combined cycle, with a Rankine as the lower cycle (in specific enthalpy) to take advantage of the emerging heat of the gas turbine (surplus, not transformable into mechanical energy).

It is well known that thermo-mechanical performance is theoretically limited to Carnot performance. Additionally, irreversibilities and frictions of all kinds that may occur throughout the cycle, which are generally defined assuming reversible processes, have an impact on performance, since it is in this area where the novelty of the cycle can be defined. What concerns the reduction of irreversibilities would be a matter of the design of the machines and equipment. Nevertheless,

image2

Given the impossibility of reducing irreversibilities to zero, it is very important that the proposed novelty is also contrasted with the real operation with irreversibilities.

An important part of the state of the art is the theoretical formulation of the Brayton cycle in its various specificities, which is briefly reviewed below:

For an open Brayton cycle, in which the air inlet to the compressor is at atmospheric pressure, and the same happens with the turbine exhaust, the ideal gas performance is

image3

image4

where 15 compression. It particularly applies to pressure-initiated isentropic compression

image5 is the thermal exponent of an isentropic evolution, where r is the ratio of

P0 and temperature T0; and reaching pressure P1 and temperature T1, the value of which is

image6

where r is the compression ratio, equal to P1 / P0.

Properly speaking, the above performance is achieved in the best case of an open Brayton, which is when the compressor and turbine outlet temperatures reach the same value, so that the Carnot ratio µ (= TM / T0) is

image7

At first glance it is strange that the performance does not depend on the extreme temperature quotient, µ, but in reality it does depend, because of the last equation, so it can be rewritten

image8

image9

This performance is always worse than Carnot's (1-µ-1). 3

image10

It is very important to note that, in the open cycle, the performance increases as does the compression ratio.

This trend changes absolutely in a closed regenerative cycle, which in the ideal gas case is governed by

image11

image12

where WC is the specific work absorbed by the compressor shaft, and WT is that delivered by the turbine shaft.

image13

image14

And therefore, in the closed Brayton of ideal gas we arrive at

image15

image16

10 It can be seen that the yield increases as µ grows, and as r decreases. Furthermore, the limit of this yield, when r tends to 1, is precisely the Carnot yield, but properly speaking, that limit does not exist, since there would be no cycle, because the high and low isobars would coincide.

15 Note in this case, that the Carnot quotient has a different expression than the previous one, since it includes the regenerator effect, which requires that the turbine outlet temperature must be higher than that of the compressor outlet. This would be written where m is the factor that characterizes the regeneration. In the equation it has been indicated that in real cases, the expansion exponents ( image17 ') and compression ( image18 ) will be different, as well as the respective compression ratios, r and r ', the latter being less than the first, for reasons of load loss in the circuit.

image19

image20

5

In the previous equation it is evident that, in each half cycle, ascending and descending in enthalpy respectively, there are three phases: compression, regeneration and external heating, in the half cycle of rising; and expansion, regeneration and external cooling in the downstream.

10

The cycle is closed by means of the entropy balance, taking into account that both compression and expansion are considered isentropic. This allows the balance to be formulated simply, as the entropy gained from the end of the compression to the beginning of the expansion by the high isobar is equal, in absolute value, to the entropy 15 lost from the end of the expansion to the beginning of the expansion. compression by the low isobar.

For each of the isobar phases, the entropy and enthalpy variation can be written in these terms, as a function of the specific heat Cp and the temperatures of the beginning and end of the phase, Txi and Txf.

image21

From this, the temperature Tx that characterizes each phase can be defined, and 20 corresponds to

image22

Where Tx0 is the temperature that would be characteristic in an ideal gas under these conditions, since in the ideal gas Cp is constant, and Cx = Gx.

image23

The entropy balance can be written, using a and b as identifiers of the high and low pressure isobars

image24

For its definition, it is necessary to identify the temperature Ta at which the

5 high pressure regeneration and external heating begins. Said Ta must be lower than the turbine outlet temperature Tt, and in fact it will be considerably lower the greater the effect of loss of enthalpy when rising through the isotherm Tc from P0 to P1, which requires considering the equation of state real, as will be seen later.

10

The definitions are thus

image25

Where the value has been entered

image26 T as the difference between hot and cold fluid in the regenerative exchanger. For the absolutely ideal case, image27 T = 0.

From the entropy balance we can write

image28

From this, and in order to link the yield with the entropy balance, the following equality can be written, taking into account that H image29 r is the same in both isobars:

image30

The yield can therefore be expressed as

image31

image32

The optimization process of the previous expression is not elementary, since all the variables depend on each other through, among other parameters, Cx and Gx which in turn condition the partition of the interval T0 to TM in each of the semicycles, from rise

5 and down. In any case, two possible situations appear, with respect to the working fluid: that it behaves as an ideal gas, or that it is a gas that differs considerably from that condition. It is well known that when gases get very close to the two-phase liquid-vapor bell, that is, to the saturation curve due to the beginning of condensation, the gas moves quite far from the ideal gas.

10

As a generalized equation of state, the Ideal Gas is used, including the so-called “compressibility factor”, identified by “z”, and that at each point is the one that makes it happen.

It is important to note that z is dimensionless, but not R (which is measured in 15 kJ / (kg · K)).

The compressibility factor z is equal to 1 when the substance behaves as an ideal gas.

There are a series of thermodynamic properties that are part of the state of the

20 art, and help formulate the invention. A useful property to characterize the regeneration phase is that the partial derivative of the specific enthalpy with respect to pressure, at constant temperature, corresponds to

image33

image34

where VT is the specific volume along the isotherm T, and fp is a parameter called "logarithmic factor of isobar expansion", and corresponds to

image35

This logarithmic factor is 0 for ideal gases.

The specific enthalpy change that occurs when passing from P0 to P1 along a given temperature isotherm, Ti; which is denoted by image36 Hi obeys

image37

where the mean value of V (VTi) and fp (fpi) have been used. The minus sign is due to the fact that the specific enthalpy decreases when the pressure increases along an isotherm.

In turn, the logarithmic isobar expansion factor is used to generally calculate the thermal exponent of temperature evolution in a real isentropic compression or expansion, which has been denoted by

image38 , which is worth, as shown below,

image39

15 and which determines the aforementioned thermal evolution, which starts in a state characterized by pressure P0 and temperature T0 and evolves to pressure P1, then the final temperature T1 is

image40

image41

To obtain the equation of

image42 It starts from the following Maxwell equation:

image43

20 in which the following equalities apply

image44

image45

It is important to remember that general isentropic evolution, characterized by dS = 0, part of the definition of entropy

image46

The real equation of state is incorporated into this equation, with z giving

image47

The previous equation indicates the logarithmic variation of the absolute temperature, with respect to the absolute pressure, in the isentropic ones, which are fundamental curves in

10 the definition of the cycle. By pointing to a pressure P0 next to T0, you are choosing an isentropic one. When the high pressure P1 is chosen, having already set the maximum admissible temperature, TM, the high isentropic that closes the cycle will be obtained.

In it there are still two fundamental things: how the regenerative heat exchange process is carried out, and how compression occurs (gas, if it is close to ideal; or steam, if that condition is sought). Additionally, there is the expansion process, but with a sufficiently high TM, this process will be very close to ideal gas. Furthermore, in expansion, the gas moves in line with its physical nature, which leads it to expand; which is just the opposite of what he suffers

in compression, which is ultimately more difficult to perform in practice, for this reason.

When considering the two aforementioned processes, compression and regeneration,

5 two enthalpic increases, measured in the low isobar, which greatly condition performance, as can be deduced from the previous equations, including the effects of the real state equation:

image48 One, the increase in enthalpy produced when passing from P0, T0 to P1, Tc; that we have denoted as

image49 Hc. 10

image50 Two, the increase in enthalpy that occurs when going down the isotherm from P1, Tc to P0, Tc. This last drop is not part of the material cycle of the fluid, but it does belong to the regenerative heat transfer process, as it is its limit, due to the extreme low enthalpy.

15 This last value of enthalpy increment, previously called image51 Hi, it is the least that is lost in the regeneration. To this must be added what is lost due to the need for the temperature of the low pressure fluid to be higher than that of the high pressure, plus the imbalances in how the specific heats vary in one and the other isobar.

20 If we take into account the specific work (specific enthalpy) of the WT turbine, the compressor, WC and the decompensated enthalpy QA (>

Image52 Hi) as a result of 100% regeneration being impossible, the performance can be written as

image53

This equation has as its numerator the definition used before for ideal gas, since in 25 the ideal gas

image54 Hi = 0, but a denominator greater than unity is added here, since the effect of the impossibility of regenerating these facilities at 100% must be included.

EXPLANATION OF THE INVENTION

The invention consists in unequivocally prescribing the essential thermodynamic variables of an assembly of components that make up a closed cycle for a working fluid whose requirements are also prescribed.

5

For this, the previous general analysis of the state of the art has to be refined, looking for assemblies between thermal devices, machines and exchangers, which offer innovation (with some type of improvement) taking advantage of the equations of state adequately.

10

It is convenient to delve into the structure of the yield equation, which with the equation of state of a real fluid, not including irreversibilities yet, could be written

image55

Where it has been distinguished, for being essential, between the exponent

image56 compression and 15

Image57 'of expansion (which will be very close to that of ideal gas, that is, R / Cp).

Regarding compression, and decompensation in regeneration, a brief collection of thermodynamic data will be relevant to this effect and will serve as a reference to the explanation, making it easier to present.

twenty

If we consider pure nitrogen, the largest component of air, we find that its critical temperature is 127 K, that is, well below that of the environment, so its behavior is very close to ideal gas, with its R = 298 J / kg K and its quotient = C Image58 p / Cv = 1.4; so the exponent

image59 = 0.286.

25

If we start from a value of T0 = 30ºC (303 K) and apply a compression of factor 2, starting from any pressure, as long as it remains as an ideal gas, it is found that the compression output Tc (reversible) is 370K; and the enthalpy consumed is 70 kJ / kg · K. If the case were closed superiorly with TM = 1000 K, the work given by the turbine would be 187 kJ / kg · K; with a turbine outlet temperature of 820 K, and a maximum theoretical efficiency of 0.63.

image60

5 If we reversibly apply a compression of ratio 2 (r = 2) to carbon dioxide, CO2, from various starting pressures, always with an initial temperature of 303 K, the following enthalpy jump values are obtained for Wc, QA and Qfn, which is the sum of the previous two, and is the minimum value to be evacuated by the cold focus (assuming everything is reversible). The theoretical maximum efficiency is added, with TM = 1000 K.

10

P0 (MPa)

Wc (kJ / (kg K) QA (kJ / kg K) Qfn (kJ / kg K) Max.

1

42 eleven 53 0.607

2

41 13 54 0.606

3

38 24 62 0.581

4

33 32 65 0.585

The previous data can be complemented with the corresponding ones of a compression ratio of factor 3 (r = 3) with the same data of maximum and minimum temperatures in the cycle. For N2 this would represent leaving the compressor at 415 K and the

15 turbine at 730 K; with respective specific works of 116.5 kJ / kg · K and 190 kJ / kg · K, which entails a performance of 0.58, less than that obtained for r = 2 (0.63).

For CO2, the data obtained for r = 3, are With these precedents, a cyclical process is presented that in its ideal definition or without irreversibilities, works between an isobar of lower pressure, or low isobar, which is at P0, and an isobar high, or higher pressure, P1, existing

P0 (MPa)

Wc (kJ / (kg K) QA (kJ / kg K) Qfn (kJ / kg K) Max.

1

68 14 82 0.598

2

63 25 93 0.567

3

60 36 96 0.575

image61

image62

a compression phase, in which a compressor draws in the fluid at its lowest specific enthalpy point of the entire cycle, at pressure P0 and temperature T0, and raises it in pressure throughout an isentropic evolution, up to P1, leaving the compressor with a temperature Tc, which is linked to T0 through which its quotient (Tc / T0) is equal to r image63 , where r is the pressure ratio (P1 / P0) and

image64

the thermal exponent of the compression;

image65

following a heating phase in which two types of heat sources act successively, which are

image66

the fluid itself, in another phase of the cycle, in which it is warmer, which is at the outlet of the turbine, thus carrying out a thermally regenerative phase, which in this heating phase reaches a temperature Ta that is below below the turbine outlet temperature, Tt, by an amount called the upper terminal temperature difference, Dtst, whose value is between 0.001 K and the temperature difference between the turbine outlet and the compressor outlet, and As reference of the invention, the value of 10 K is taken;

image67

the external source of heat input to the working fluid, with which the fluid is heated up to TM, which is the maximum temperature that the working fluid reaches, selecting the source of said heat from the combustion of a fuel in a chamber combustion outside the closed circuit of the working fluid, or another source of heat such as solar thermal, transferring the heat generated to the working fluid, through a heat exchanger called a heater;

Image68

an expansion phase, from the point of maximum specific enthalpy of the cycle, in which the working fluid is at pressure P1 and temperature TM, evolving isentropically up to pressure P0, leaving the turbine

or expander machine, where this phase is carried out, with a temperature Tt;

image69

following a cooling phase in which two types of cooling actions concur, which are

5

10

fifteen

twenty

25

30

image70

the fluid itself, in another phase of the cycle in which it is colder, which is at the compressor outlet, and which is carried out in a regenerative heat exchanger, cooling the fluid to a temperature Tb that is above the temperature Tc of compressor output, in an amount called the lower terminal temperature difference, represented by Dtit, whose value is between 0.001 K and the temperature difference between the turbine outlet and the compressor outlet, and as a reference of the invention, the value of 10 K;

image71

the external cooling sump, which cools the fluid down to T0

and in which the fundamental prescriptions are

image72

the quotient between the maximum and minimum temperatures of the fluid, TM / T0, must be greater than the compression ratio r, corresponding to the quotient P1 / P0, raised to the sum of the thermal exponents in the isentropic evolutions of compression and expansion, which is equivalent to establishing that the turbine outlet temperature, Tt, is greater than the compressor outlet temperature, Tc;

image73

the pressures of the low isobars, P0, and high, P1, are set by meeting the criteria:

-that the specific enthalpy gained by the working fluid, when descending through the compressor output isotherm, from P1 to P0, in a virtual process that is not part of the real cycle, is less than the specific work of the compressor, which is which is specified as

image74

where fp0 is the logarithmic factor of isobar expansion of the working fluid for the compressor outlet temperature and a pressure that is the geometric mean of P0 and P1, and Vc0 is the specific volume of the fluid at said compressor outlet temperature, and said geometric mean of pressures;

and for the pressures P0 and P1, the isobar specific heats of the working fluid satisfy the following equality, with a tolerance of ± 20%, being

image75

image76

image77

image78

image79 a temperature difference whose value can be selected between 0.1 ° C and 100 ° C, with 10 ° C being the value considered as a reference in this invention.

5 To determine the optimum value of the compression ratio, r, the value of the regeneration factor that produces this optimum, m, is found from the Carnot quotient of the case, µ, which is factored into three factors, corresponding to the compression, regeneration and expansion phases, according to the following equation, which differentiates between the ratio of

10 compression, r, in the compressor, and r ', in the turbine,

image80

where r '<r for reasons of manometric head loss throughout the entire working fluid circuit, which is fundamentally a function of the value of the regeneration factor, m, since the higher the value of the enthalpy to be regenerated , greater pressure drop, which is valued in this invention by means of a reduction coefficient,

image81 , such that r '=

image82 R bearing

image83 the following dependence on m, as investigated in this invention

image84

from which the simplified performance equation is modified, without accounting for the pressure drop, which is

to get the realistic performance

image85

image86

that it has to be maximized with respect to m, which is the independent variable here, adopting for m the value that maximizes the realistic yield, with a tolerance of ± 20% in m; being able to make a numerical approximation to determine said value of m, as shown below for Argon, with a value of

image87 = 0.40; and a case of µ = 3.3333, typical of T0 = 300 K and TM = 1000 K, which is presented in the table below,

image88

Performance Simplified realistic m r '/ r performance

1 1 0.4522747 0.4522747 1.1 0.9977348 0.47712384 0.47776442 1.2 0.99094946 0.49724995 0.4999975 1.3 0.97967472 0.512932 0.51961314 1.4 0.96396167 0.5241268 0.53708764 1.5 0.94388148 0.53041946 0.55278417 1.6 0.91952513 0.53087379 0.56698513 1.7 0.89100297 0.52369449 0.57991387 1.8 0.8584442 0, 50546332 0.59174967 1.9 0.82199634 0.46924209 0.60263831

2 0.7818245 0.399027 0.61269973

where it can be seen that the simplified yield is monotonically increasing with m, while the realistic yield presents a maximum for m = 1.6; which implies a reducing coefficient ( image89 ) of 0.919, and taking into account that

image90

image91

and with the given values, it is obtained that the compression ratio is r = 2.6.

The embodiment of the invention is selected between making it with constant content

10 over time, without varying the mass of fluid contained in the closed circuit of the system, or with a variable content of working fluid in the circuit, depending on the operating conditions; and in case of selecting variable fluid inventory during operation, there is a low pressure fluid extraction valve, before the compressor suction, and a fluid injection, controllable with a valve,

15 connected to a tank with a pressure higher than that of the circuit at that point.

The working fluid is selected from a pure substance or a mixture of substances, chosen in such a way that its specific heat at constant pressure, at the pressures and temperatures of interest, better meets the requirements that Cx be less than Gx in the process of compression, that Cx of the low isobar, in the regeneration process, at a temperature T, is equal to the Cx of the high isobar at a temperature T- image92 T, being

image93 T is a temperature difference whose value can be selected between 0.1 ºC and 100 ºC, with 10 ºC being the value considered as a reference, and the aforementioned equality being fulfilled with a tolerance of ± 20%.

The compressor drive is selected between being carried out mechanically from the turbine shaft, or activated by an electric motor, whose power supply can be selected from various sources.

In the case of keeping constant the content of the working fluid inside the device, its operation is regulated, at least, by the contribution of heat through the heater, or hot source, and the extraction of heat through the cold source. ; and in the event that the compressor drive by electric motor is selected, the system is also regulated by the power given to the compressor, and its revolutions.

EXPLANATION OF THE FIGURES

Figure 1 shows a diagram of an assembly of a device in which the process of the invention could be implemented.

Figure 2 shows the lines of the processes and the main thermodynamic points of the cycle, in a graph (enthalpy, log P).

To improve understanding of the explanation of the figures, the elements that make up the invention are listed below:

1.

Gas turbine)

2.

Turbine exhaust and connection to the low pressure circuit of the regenerative exchanger

3.

Low pressure circuit of the regenerative exchanger

Four.

Regenerative exchanger

5.

Low pressure circuit outlet of the regenerative exchanger

6.

Working fluid circuit in heat sink

7.

Heat sink

8.

Outlet of the working fluid from the heat sink, and connection to the compressor inlet

9.

Outside heat sink refrigerant inlet

10.

Outside refrigerant outlet from heat sink

image94

5 11. Compressor

12.

Compressor outlet and connection to the high pressure circuit of the regenerative exchanger

13.

Compressor Drive Electric Motor (11)

14.

Electric generator activated by the turbine shaft (1)

10 15. Countercurrent circuit of the working fluid in the cold branch, high pressure, in the regenerator.

16. Outlet of the high pressure circuit of the regenerative exchanger and connection to the hot bulb

17.

Hot spot, which may be constituted by the fumes of atmospheric combustion, or from another source

18.

Outlet, from the high temperature source, of the fluid at high pressure and at the highest temperature, for entry into the turbine

19.

Atmospheric combustion chamber

twenty.

Fuel injector, or burner 20 21. Intake of preheated atmospheric air into chamber (19)

22.

Conduction of hot fumes from the chamber (19) to the exchanger that acts as a hot spot (17)

2. 3.

Conduction of the fumes from its exit from the hot spot (17) to the

heat recovery unit (24) 25 24. Air / air heat exchanger, or recuperator

25.

Inlet of atmospheric air into the recuperator

26.

Smoke evacuation chimney

27.

Thermodynamic point of minimum enthalpy of the cycle, which is the entrance to the

compressor (11). 30 28. Compressor outlet point (11).

29.

Point at which the regeneration in the high isobar ends, and the fluid enters the heater, its temperature is Ta which is equal to Tt minus Dtst.

30.

Entry point into the turbine (1).

31. Turbine outlet point. 35 32. End point of regeneration in the low isobar (its temperature

is Tb = Tc + Dtit) 18

image95

33.

Temperature point Tc in the low isobar

3. 4.

Point, in the low isobar, with the same enthalpy as that of the compressor outlet (28).

In the diagram in figure 2 there are several lines defined by a property: the LP0 mark indicates the line in which the pressure is constant and is equal to P0. LP1 is the high isobar.

LTt designates the isotherm of the turbine exhaust temperature, and LTc the isotherm of the drive from the compressor.

T0 is the temperature of the point of minimum enthalpy of the cycle (27), and TM that of maximum enthalpy (30).

LSc is the isentropic compression.

LSt is the isentropic expansion in the turbine.

Lri represents the conditions of the regenerator on the lower face (enthalpy) and Lrs the conditions of the upper face.

LE is the isenthalpic line that goes from point 28 to 34.

MODE OF EMBODIMENT OF THE INVENTION

The invention is materialized by appropriately integrating the two physical elements that compose it: the working fluid, and the thermo-mechanical equipment in which the processes that make up the cycle take place.

For this, a set of thermal engineering elements or components is available, ranging from the compressor (11) to the turbine (1), passing through the regenerative exchanger (4), plus the heater or hot spot (17) to heating the high pressure fluid with the external means (19). Before the compressor, in the succession of phases of the cycle, there is the cold bulb (7).

image96

5 Materialization includes as an essential issue setting the levels of the relevant variables in the definition of the cycle, complying with the prescriptions already expressed. These levels are substantially dependent on the working fluid used.

Figure 1 shows a diagram that displays a preferred embodiment of

10 the invention. The device on which the thermodynamic cycle is carried out comprises a gas turbine (1), whose exhaust (2) is connected to a low pressure circuit (3) of a regenerative exchanger (4) which, as seen below, it also has a high pressure circuit. Thermodynamically, the points necessary to define each process are given in figure 2, with the expansion in the turbine going from point 30 to

15 31, and regenerating in the lower isobar from point 31 to 32.

The outlet (5) of said low pressure circuit (3) is connected to an external heat sink (7) through the corresponding circuit (6) of the working fluid in said sink. This heat sink (7) is cooled, having an inlet (9) and

20 an outlet (10) of the external coolant from the heat sink. The outlet (8) of the working fluid from the heat sink (7) sets the conditions in which said working fluid enters the compressor (11).

Once the working fluid has finished its evolution in the compressor (11), the

The outlet (12) of the compressor (11) involves the connection to the high pressure circuit (15) of the regenerative exchanger (4). Thus, when the working fluid emerges from the outlet

(16) of the high pressure circuit of the regenerative exchanger (4), is in a position to start its evolution in the hot spot (17), increasing its enthalpy. At the outlet (18) of said hot spot (17), the fluid is in conditions of

30 evolve in the turbine (1), thus initiating a new cycle.

image97

The main characteristics of the thermodynamic cycle characteristic of this invention are shown in figure 2, whose thermodynamic points are directly identified with geometric and physical points of the circuits and machines of the device.

A novelty in the invention is to keep the total fluid content fixed, which is distributed throughout the circuit according to temperature and pressure.

It should be noted that in normal operation the situation within each section of the circuit will be almost isobar, because if there were considerable pressure differences, the fluid would accelerate enormously. The pressure in question can vary over time, as the enthalpy load and the T vary, but this evolution can be considered as a succession of cases almost in equilibrium.

At rest (without heating or compression) the system will be isothermal and isobaric, and the properties are valid for all points.

In operation, the system will have to distinguish essentially four sections: the high circuit, the turbine, the low circuit, and the compressor.

Each of them will have to fulfill:

image98 The equation of state, at each thermodynamic point, and in the total of the section, with average values in the intensive variables

image99 The mass continuity equation

Regarding any section, generically called i, you can write:

image100

where Vi is the volume of each section. In turn, each section can be subdivided into consecutive compartments, as required to refine in the description of P and T, and as averages in machines, where there is strong variation of P and T, the geometric mean of each variable can be taken , between the inlet and outlet of the compartment.

Regarding the continuity equation, we can write M'i = constant, where M 'is the mass flow. To determine what fluid content, M, is given in total to meet the nominal specifications, one can write, for each section:

image101

where image102 i is the time of passage through section i, and the above equation can be applied to the four sections, a, high, t turbine, b low, and c compressor

image103

Where M is the total mass of inventory to meet the nominal specifications, and

image104 the total travel time of the cycle

image105

where V is the total volume of the system.

When the system is in stop conditions, at a uniform temperature T ’, the pressure P’ that it will reach will be

Image106

Claims (1)

image 1 1 - Thermodynamic cyclical process with turbine and gas compressor, with heat input from an external source, where the cycle works between a lower pressure isobar, or low isobar, which is at P0, and a high isobar, or higher pressure, P1, existing:

image2

a compression phase, in which a compressor draws in the fluid at its lowest specific enthalpy point of the entire cycle, at pressure P0 and temperature T0, and raises it in pressure throughout an isentropic evolution, up to P1, leaving the compressor with a temperature Tc, which is linked to T0 through which its quotient (Tc / T0) is equal to r image3 , where r is the pressure ratio (P1 / P0) and

image4

the thermal exponent of the compression;

image5

following a heating phase in which two types of heat sources act successively, which are

image6

the fluid itself, in another phase of the cycle, in which it is warmer, which is at the outlet of the turbine, thus carrying out a thermally regenerative phase, which in this heating phase reaches a temperature Ta that is below below the turbine outlet temperature, Tt, by an amount called the upper terminal temperature difference, Dtst, whose value is between 0.001 K and the temperature difference between the turbine outlet and the compressor outlet, and As reference of the invention, the value of 10 K is taken;

image7

the external source of heat input to the working fluid, with which the fluid is heated up to TM, which is the maximum temperature that the working fluid reaches, selecting the source of said heat from the combustion of a fuel in a chamber combustion outside the closed circuit of the working fluid, or another source of heat such as solar thermal, transferring the heat generated to the working fluid, through a heat exchanger called a heater;

image8

an expansion phase, from the point of maximum specific enthalpy of the cycle, in which the working fluid is at pressure P1 and temperature TM, evolving isentropically up to pressure P0, leaving the turbine

or expander machine, where this phase is carried out, with a temperature Tt;

image9

following a cooling phase in which two types of cooling actions concur, which are

image10 image11 the fluid itself, in another phase of the cycle in which it is colder, which is at the compressor outlet, and which is carried out in a regenerative heat exchanger, cooling the fluid to a temperature Tb that is above the temperature Tc compressor outlet, in a quantity called the lower terminal temperature difference, represented by Dtit, whose value is between 0.001 K and the temperature difference between the turbine outlet and the outlet of the 10 compressor, and as a reference of the invention the value of 10 K is taken; the external cooling sump, which cools the fluid down to T0 image12 characterized in that the prescriptions that the cycle variables meet are: image13 the quotient between the maximum and minimum temperatures of the fluid, TM / T0, must 15 be greater than the compression ratio r, corresponding to the ratio P1 / P0, raised to the sum of the thermal exponents in the isentropic evolutions of compression and expansion, which is equivalent to establishing that the turbine outlet temperature, Tt is greater than the compressor outlet temperature, Tc; 20 the pressures of the low, P0 and high isobars P1 are fixed because the specific enthalpy gained by the working fluid, when descending through the compressor outlet isotherm, from P1 to P0, in a virtual process that is not part actual cycle time, is less than the specific work of the compressor, which is specified as image14 image15 25 where fp0 is the logarithmic factor of isobar expansion of the working fluid for the compressor outlet temperature and a pressure that is the geometric mean of P0 and P1, and Vc0 is the specific volume of the fluid at said compressor outlet temperature, already said geometric mean of pressures. 30 2 - Thermodynamic cyclical process with turbine and gas compressor, with heat input from an external source, according to claim 1, characterized in that for the image16 image17 image18 being image19 a temperature difference whose value can be selected between 0.1 ° C and 100 ° C, with 10 ° C being the value considered as a reference in this invention. 5 3 - Thermodynamic cyclical process with turbine and gas compressor, with heat input from an external source, according to the first or second claim, characterized in that the optimal value of the compression ratio is determined by finding the value of the regeneration factor, m, from the Carnot quotient of the case, µ, which is factored 10 in three factors, corresponding to the compression, regeneration and expansion phases, according to the following equation, which differentiates between the compression ratio, r, in the compressor, and r ', in the turbine, image20 where r '<r for reasons of manometric head loss throughout the entire working fluid circuit, which is fundamentally a function of the value of the regeneration factor, m, since the higher the value of the enthalpy to be regenerated , greater pressure drop, which is valued in this invention by means of a reduction coefficient, image21 , such that r '= image22 R bearing image23 the following dependency with m, image24 image25 from which the simplified performance equation is modified, without accounting for the pressure drop, and assuming ideal gas, which is to get the realistic performance image26 image27 that has to be maximized with respect to m, which is the independent variable, adopting for m the value that maximizes the realistic yield, with a tolerance of 25 ± 20% in m. image28 4 - Thermodynamic cyclical process with turbine and gas compressor, with heat input from an external source, according to previous claims, characterized in that the working fluid is selected from a pure substance or a mixture of substances, chosen in such a way that its specific heat is constant pressure, at the pressures and temperatures of interest, meet the requirements that Cx be less than Gx in the compression process; and that Cx of the low isobar, in the regeneration process, at a temperature T, is equal to the Cx of the high isobar at a temperature T- image29 T, being image30 T is a temperature difference whose value can be selected between 0.1 ºC and 100 ºC, with 10 ºC being the value considered as a reference, and the aforementioned equality being fulfilled with a tolerance of ± 20%. 5 - Thermodynamic cyclical process with turbine and gas compressor, with heat input from an external source, according to claim one, characterized in that the system works with constant total fluid content, and in operation, in the system it will be necessary to distinguish at least four sections: the high circuit, the turbine, the low circuit, and the compressor, and each of them must comply with: image31 the equation of state, at each thermodynamic point, and in the total of the section, with average values in the intensive variables, and image32 the mass continuity equation; and with respect to any section, generically called i, you can write: image33 where Vi is the volume of each section; and in turn, each section can be subdivided into consecutive compartments, as required to refine in the description of P and T, and as averages in machines, where there is strong variation of P and T, the geometric mean of each variable can be taken , between the inlet and outlet of the compartment; and for what corresponds to the continuity equation, we can write M’i = constant, where M ’is the mass flow, being fulfilled in each section: image34 where image35 i is the time of passage through section i, and the above equation can be applied to the four sections, a, high, turbine t, low b, and compressor c, so that image36 image37 where M is the total mass of inventory to meet the nominal specifications, and image38 the total travel time of the cycle image39 where V is the total volume of the system; and when the system is in shutdown conditions, at a uniform temperature T ', the pressure P' that it will reach will be image40 6 - Device for carrying out a thermodynamic cyclic process with a gas turbine and compressor, with heat input from an external source, according to any of the first to fifth claims, characterized in that said device on the 10 that the thermodynamic cycle is carried out comprises a succession of equipment in which processes take place along which, and over time, the total content of working fluid is kept constant, if the inventory option of fixed fluid, the equipment being: - a gas turbine (1), the exhaust (2) of which is connected to the inlet of the low pressure circuit (3) of the regenerative exchanger (4); - a regenerative heat exchanger (4), which integrates a section of the low pressure circuit (3), of the circuit that runs through the working fluid, and in countercurrent a section of the high pressure circuit; and in this high pressure circuit, the heating phase in this regenerator reaches a temperature Ta 20 that is below the turbine outlet temperature, Tt, by an amount called the upper terminal temperature difference, Dtst, whose value is between 0.001 K and the temperature difference between the turbine outlet and the outlet of the compressor, and as a reference of the invention the value of 10 K is taken; cooling the fluid in the low pressure circuit to a temperature Tb that is above 25 the compressor outlet temperature Tc, in a quantity called the lower terminal temperature difference, represented by Dtit, whose value is between 0.001 K and the temperature difference between the turbine outlet and the compressor outlet, and as a reference of the invention takes the value of 10 K; image41 -a heat sink (7) connected to the outlet (5) of the low pressure circuit (3) of the regenerative exchanger (4), said heat sink (7) being cooled, having an inlet (9) and an outlet (10) for external refrigerant;

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a compressor (11), whose inlet is connected to the outlet (8) of the working fluid of the heat sink (7), and whose compressor outlet is connected to the inlet of the high pressure circuit of the regenerative heat exchanger (4 );

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a hot bulb (17), thermally fed from the outside, and into which the working fluid enters, in its high pressure circuit, from the regenerative heat exchanger (4), and whose outlet (18) is connected to the inlet to the gas turbine (1).

7 - Device for carrying out a thermodynamic cyclical process with a turbine and gas compressor, with heat input from an external source, according to claim six, characterized in that the embodiment of the invention is selected between making it with constant content throughout the time, without varying the mass of fluid contained in the closed circuit of the system, or with a variable content of working fluid in the circuit, depending on the operating conditions; and in case of selecting variable fluid inventory during operation, there is a low pressure fluid extraction valve, before the compressor suction, and a fluid injection, controllable with a valve, connected to a higher pressure tank to that of the circuit at that point. 8 - Device for carrying out a thermodynamic cyclical process with a gas turbine and compressor, with heat input from an external source, according to any of the sixth or seventh claims, characterized in that the compressor drive is selected between being mechanically performed from the shaft of the turbine, or activated by an electric motor, whose power supply can be selected from various sources. 9 - Device for carrying out a thermodynamic cyclical process without condensation of the fluid and with limited prescriptions on its minimum and image42 according to any of the sixth or seventh, and eighth claims, characterized in that in the case of keeping constant the content of the working fluid inside the device, its operation is regulated, at least, by the contribution of heat through the heater, or hot bulb, and the extraction of heat through the cold bulb; and in the event that the compressor drive by electric motor is selected, the system is also regulated by the power given to the compressor, and its revolutions.