THERMODYNAMIC CYCLIC PROCESS WITHOUT FLUID CONDENSATION AND WITH PRESCRIPTIONS TAXED ON ITS POINTS O

专利摘要:
Thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its points of minimum and maximum enthalpy, and device for its realization. A gas cycle is prescribed in which the temperature of the minimum enthalpy point of the cycle is below the critical temperature of the working fluid, and above the ambient cooling temperature; and the pressure of that point is above the critical pressure. This pressure of the low isobar branch is chosen as a function of a parameter that is denoted as a logarithmic factor of isobaric dilatation, with which properties are established between the isobars and isentropics of the cycle. To carry out the process, it is necessary to use a source and auxiliary heating flow, in the regenerative exchanger, either continuously or in jumps. (Machine-translation by Google Translate, not legally binding)
公开号:ES2652522A1
申请号:ES201731263
申请日:2017-10-30
公开日:2018-02-02
发明作者:Jose Maria Martinez-Val Peñalosa；Javier Muñoz Anton；Luis Francisco GONZALEZ PORTILLO；Ruben Amengual Matas
申请人:Universidad Politecnica de Madrid；
IPC主号:

专利说明:

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THERMODYNAMIC CYCLIC PROCESS WITHOUT CONDENSATION OF THE FLUID AND WITH PRESCRIBED LIMITS ON ITS MINIMUM AND MAXIMUM ENTALPY POINTS, AND DEVICE FOR THEIR REALIZATION
DESCRIPTION
SECTOR OF THE TECHNIQUE
The invention falls within the field of thermodynamic cycles that transform thermal energy into kinetic energy of rotation, in a turbine where the working gas is expanded.
TECHNICAL PROBLEM TO BE RESOLVED AND BACKGROUND OF THE INVENTION
The problem is to maximize the performance of the installation where the cyclic process is carried out. For this, the properties of the fluids in the different domains they manifest, characterized by certain ranges of pressure and temperature, must be studied rigorously.
Note that the pressure is limited by the thickness given to the walls of containers and pipes, and by the quality of the material that constitutes them, while the temperature is generally limited by the physical process where heat is generated, that is, where it is it gives the fluid (which in our case will be in a supercritical state) the maximum possible enthalpy, and is the highest of the entire cycle that the fluid experiences.
It is well known that thermo-mechanical performance is limited to Carnot's performance, theoretically. In reality, the 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 that area where the novelty of the cycle can be defined. What concerns the reduction of irreversibilities would be a matter of machine and equipment design.
As a direct precedent, which in turn has in its exhibition the analysis of various backgrounds of these cycles, it is necessary to mention the patent ES 2427648 B2, which deals with a cycle
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Brayton with environmental cooling close to the critical isotherm, its first inventor being the same as that of this application.
The invention presented here not only improves the exploitation of the physical characteristics of the usual or potential working fluids, but also seeks that the processes can be carried out in the machines in the best way. The result is a Brayton type cycle, therefore without condensation, with very precise specifications in the definition of the minimum and maximum enthalpy points of the working fluid that executes the cycle.
The most important antecedents of the invention are the thermodynamic relationships themselves established as universal rules, but which materialize in very specific physical behaviors as one or the other alternative is taken in the specifications to be defined.
EXPLANATION OF THE INVENTION
The invention consists in specifying unambiguously defined prescriptions, on a thermodynamic cyclic process, which in its ideal definition or without irreversibilities, works between an isobar of lower pressure, or low isobar, which is at P0, and a high isobar, or of greater pressure, P1, existing
- a compression phase, in which a compressor aspirates the fluid at its point of least specific enthalpy of the entire cycle, at pressure P0 and temperature T0, and elevates it from pressure along an isentropic evolution, up to P1, leaving of the compressor with a temperature Tc;
- following a heating phase in which three types of heat sources concur, which are
or the fluid itself, in another phase of the cycle, in which it is hotter or the main source of heat input to the working fluid, with which the fluid is heated to TM, which is the maximum temperature reached by the fluid of job
or an auxiliary thermal source, which can act on the working fluid either at the same time as the main source, either alternately with it, but without reaching the TM temperature, in any case;
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- 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 the pressure P0, leaving the turbine or expanding machine where this phase is performed with a temperature Tt;
- following a cooling phase in which two types of cooling actions concur, which are
or the fluid itself, in another phase of the cycle in which it is colder or the external cooling sump, which cools the fluid to T0
and in which the fundamental prescriptions are
- the pressure of the specific minimum enthalpy point, P0, is greater than the critical pressure, Pcr, of the working fluid; and its temperature, T0, is lower than the critical temperature, Tcr, of said fluid;
- the ratio between the maximum and minimum fluid temperatures, TM / T0, must be greater than the compression ratio r, corresponding to the ratio P1 / P0, raised to the sum of the thermal exponents in isentropic compression and expansion evolutions , which is equivalent to establishing that the turbine outlet temperature, Tt, is higher than the compressor outlet temperature, Tc;
The prescriptions are completed with the properties or relationships established in what
follow, between various parameters of the working fluid state equation.
In principle, it will be a non-degradable fluid in the ranges of variables to be assumed, and whose state can be determined with two variables, such as temperature and pressure. On them the Volume, Enthalpy and Entropy will depend, through the equation of state, all of them in their specific formulation, that is, per unit mass, thus being the respective units m3 / kg; kJ / kg; and kJ / (kgK).
As a generalized state equation, the Ideal Gas will be used, including the so-called "compressibility factor", identified by "z", and which at each point is the one that makes compliance
P-V = z-R-T
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It is important to note that z is dimensionless, but not R (which will be measured in kJ / (kgK)). For applications, the International System of Units should be used, but this is not part of the invention or its explanations.
The compressibility factor z is worth 1 when the substance behaves as an ideal gas; but in general it differs quite a lot from that value.
An essential thermodynamic point in any substance is the critical point, defined by its critical temperature and its critical pressure, and its corresponding critical volume exists. It is the point of maximum pressure and maximum temperature of the curve that delimits inside the biphasic curve of change of liquid-vapor state. Above that temperature, the gas cannot condense. And in addition there is no discontinuity in the specific volume of the fluid above the critical pressure, nor is there conventional boiling, with constant temperature and pressure while the phase change occurs.
If for a substance the values of P, T and V are determined experimentally at the critical point, denoted as Pcr, Tcr and Vcr, its compressibility factor is found at the critical point, zcr, which is
P V
1 cr vcr RT, and
image 1
Said zcr is worth around 0.25 for any substance (in fact, the vast majority of known substances have their zcr between 0.2 and 0.3), but it is even more important to analyze the topology of the z values, which can be explained with a remarkable approximation depending on the specific enthalpy of the fluid, represented by H and measured in kJ / kg.
There is a domain in the thermodynamic diagram of the substances of interest, which extends to enthalpies smaller than that of the critical isotherm, for pressure values above Pcr; which logically can be called "supercritical pressure domain, low enthalpy", or DSPBE. In it, the value of z depends fundamentally on pressure, and almost nothing on T and V. In fact, it can be written for this domain:
image2
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In this expression w is an exponent somewhat smaller than the unit (0.9 for the substances of interest) and PR is a reference pressure, which can be set at 25 MPa, and that does not really affect the complete thermodynamic analysis at all and its conclusions; Since the defined cycle is between two isobars, high and low, and both referred to PR, the overall effect of it is canceled. Not so that of the exponent w, which is very important to characterize thermodynamic transformations in this domain.
In the thermodynamic approach that is needed, the value of the z parameter is not sufficient as a point function to make two energy terms equal, (PV) and (RT); but it is necessary to have its most important variations characterized, for which a parameter, fp, not previously used in Thermodynamics (according to the authors' knowledge) is introduced, but which has its utility to calculate the partial derivative of the specific enthalpy with respect to the pressure, at a constant temperature, which corresponds to
image3
image4
VT being the specific volume along the T isotherm, and fp a parameter that has been called "logarithmic factor of isobar dilation", and corresponds to
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p
This parameter has great utility to define a line in the thermodynamic diagram, mainly in which specific enthalpy is used in abscissa and pressure (in logarithmic scale) in ordinates, or diagram (H, logP) and is a proposed line, in the opinion of the inventors, called "secondary discontinuity line", and is the geometric place that form the points of highest value of fp in each isobar, above the critical pressure. That line delimits two different regions in the mentioned diagram. left of it, the curves of z = constant, or iso-level curves of z, are almost horizontal in the above diagram (H, logP); while to the right of the line, as the enthalpy increases, the curves iso-level of z acquires a vertical component, which becomes dominant, and even more, in that line of secondary discontinuity, or in its vicinity, the iso-level curves of z with a value less than approximately 0.5 disappear.
In turn, this parameter serves to calculate in general the exponent of temperature evolution in an isentropic compression or expansion, which has been denoted by g, which is worth, as shown below,
zR
9 = t (1 + fp)
and which determines the aforementioned thermal evolution, which starts in a state characterized by 5 the pressure P0 and the temperature T0 and evolves to the pressure P1, fulfilling then that the final temperature T1 is
I ±
T0
image6
to
It should be noted that the following values of these parameters are available for an ideal gas
10 z = 1
fp = 0
g = (Y-1) / Y with Y = Cp / Cv
Cp and Cv being the specific heats at constant pressure and at constant volume.
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To obtain the equation of g, we start from the following Maxwell equation:
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p
in which the following equalities apply
image8
And by combining both derivatives the previous value of g is obtained:
P (dT ^ fdLnT ^
T dp) s ~ dLnp) s
image9
In the supercritical pressure domain, low enthalpy (DSPBE) the exponent g is much smaller than 0.1; while for high temperature values (double the criticism or more) g is worth over 0.1; being close to 0.2 for heavy molecules, and 0.3 for 5 the lightest (such as air).
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As a rigorous theoretical support, the analysis of isentropic evolution, characterized by dS = 0, is added, where S is the entropy
dT dP
dS = Cv - + (-) irdV
The equation of state including z is incorporated into the analysis, which gives
fdP R ((dz
(dz fdz idd fdz R ((dz V
df) v = dp) v df) v = dp) v [V V df) v + J
(~)
dT) v
R
pw ^
WK ~ z
R
Z
Wz
~
(- W Z ■
dTjv and v 1 -wT + Z) 1
R
KdTjv and V ‘i -wT 'J 1-wV being able to use this last equation in the isentropic condition
dT
dS - 0 - Cv —h
R
1-wV
■ dV
From which the relationships between the variations of P, T and V are deduced
dT
C J T7 "-
zR dV
T 1-w V zR dV zR dP
dT
~ T ~ ~ (1 - w) Cv ~ ~ ~ C ^ ~ P dV (1 - w) Cv dP ~ V ~
y0 pj
p
P, '~ m
m being the exponent of logarithmic density variation,
(1 - w) Cv
m = ■
which bears the sign - in the exponent when applied to the inverse of the density, which is the specific volume.
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For the ideal gas case, z is worth 1, and therefore does not depend on P, which is equivalent to w = 0. If it is applied to the last equation, m = 1 / Y remains (as is already known).
On the contrary, in the DSPBE w is worth almost 1, and m is a tiny value, which is essential for compression in the DSPBE zone.
For this, it should be added that the derivative of the specific enthalpy with respect to pressure, at constant entropy, is the specific volume (VS, along the isentropic, given by a constant value of the entropy S). This is
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This allows the specific compression work to be calculated to pass from the state of minimum specific enthalpy in the fluid, characterized by P0 and T0, at the inlet of the compressor, and bring it under pressure P1 and the corresponding T1, which will be
Ti = T0rgc
R being the compression ratio, which is equal to P1 / P0; and gc being the exponent g 15 particularized for isentropic compression. The work in question is expressed in terms of specific enthalpy variation, and is denoted by AHC, corresponding to
v0
A ^ = TT ^ (pirSC “po)
1 + gc
Similarly, the work done by the expander or turbine machine, AHe, is calculated, although in this case the really binding or limiting variable is the maximum temperature 20 attainable by the working fluid with the conditions of the hot spot available , denoting said temperature by TM.
... _7p / r
AHe = —------- (1 - r zK / lp)
where z and Cp are the average values of said coefficients along the isentropic in question, and zM the value of z corresponding to the point of maximum enthalpy.
It remains to formulate the specific enthalpy variation that occurs when passing from P0 to P1 along a given temperature isotherm, T¡; which is denoted by AH¡ and obeys
VfvdP = -VTifpi (P, - P0)
where the average value of V (VTi) and fp (fp) has been used. The minus sign is due to the specific enthalpy decreases when the pressure is increased along an isotherm.
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It should be added that in the thermodynamic window that represents a cycle of this type, which is delimited by two values, minimum and maximum pressure, which are P0 and P1, and by two extreme lower and higher temperatures, T0 and TM, there is a intermediate temperature, denoted by Ta, in which the following asymmetry is observed: for temperatures higher than 10 Ta, the specific isobar heat is higher at higher pressure; and at temperatures below Ta, the specific isobar heat is higher at lower pressure, which means that in Ta a strangulation appears in the heat transfer process in the regenerative heat exchanger, which requires finding suitable solutions to the configuration and appropriate operating conditions of said exchanger.
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The latter is associated with the process of enthalpy variation along an isobar, which in turn has repercussions on S, T and V, which can be characterized from the equation
P _zT R ~~
In it, the member of the left remains constant in an isobar, so it follows, taking differentials
dz dT 0 = - + - ■ z T
dV
~ V
Clearing dV / V and dividing all of it by dT / T you have
image13
- fP + 1
And dividing by the specific heat at constant pressure, Cp, is obtained by entering the entropy variable, S:
image14
This relationship is very important, since the cycle is delimited by two isentropics, 5 compression and expansion, which means that the increase in entropy when the fluid is heated by the high isobar is equal in absolute value, in the reversible approximation followed, to the reduction of entropy by the low isóbara, and so much by one as by another the previous relation is fulfilled.
10 If the specific volumes are considered at the beginning and at the end of the compression, denoted by Vc0 and Vc1, and the specific volumes at the beginning and at the end of the turbine expansion are also denoted as Vt1 and Vt0, you can write
image15
The importance of the quotient (fp +1) / Cp is thus appreciated in order to obtain a considerable growth of V for a given increase in entropy.
In addition, in the isentropic compression the variations of T and V are lowercase, taking the form
image16
V,
he
V,
co
Where the exponents are very small in absolute value, typically gc = 0.02 (and less than 20 than 0.05, due to specifications restriction) and m = 0.05 (and less than 0.1 for the same reason).
If we now consider how V grows along an isobar, as its temperature increases, and therefore its entropy, they are called A0 and A1
image17
And, for the specific volumes at the entrance and the exit of the turbine, the ratio is obtained through the behavior in the high and low isobars, the ratio
v,
_P_ _ r m exp (^ ¿4l _ j40)
v,
t o
5 On the other hand, in the expansion isentropic you have
TJL = rat
Tt0
This leads to being able to write
Vtl ZM
Vt0 zt0
Where the exponent gt will be very close to (y - l) /y.For values of TM above twice the Tcr (in absolute scale) which are those that have an interest in these applications, zM and zt0 are practically equal, of so that its quotient can be taken as 1.
10 By matching this last equation to the previously written ratio for those same volumes, the closing ratio of the cycle is obtained
image18
■ ^ o Ai
1 - gt - m
More importantly for performance characterization is the comparison, in absolute value, of the loss of specific enthalpy when rising through the isotherm of the turbine outlet temperature, Tt0, from the pressure P0 to P1, with the enthalpy converted in specific work of the turbine, it is convenient that the first one be small with respect to the latter, to obtain a good performance; since, at a minimum, the main source of heat will have to supply the sum of the absolute values of both quantities, of which only the second of the quantities becomes useful work. If fpt0 is the logarithmic factor 20 of isobar dilation, averaged over the mentioned isotherm of Tt0, and Vt0 is the
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specific volume at the output of the turbine, the absolute value of its specific enthalpy variation, AHt0, is
AHt0 = fPtoVt0P0Ln (r)
And the specific enthalpy turned into specific work of the turbine, AHtt, is (assuming that the value of z is already very close to 1 in said isentropic)
gt
To have a good performance, the enthalpy ceded in the AHtt turbine must be greater than AHt0, which means that fpt0 is bounded
yQt - ^
fpto <Ln (r)
It is important to note that for usual values of r in this discipline, between 1.5 and 5, and also usual values in gt, below 0.3; the value obtained in the ratio of the right member of this last equation is always slightly greater than 1. For example, for r = 3 and gt = 0.1 the value of fpt0 * is 1,057; and goes up to 1,118 for gt = 0.2. With this last exponent and r = 4 the value 1,152 comes out.
There is an isothermal line of importance, which corresponds to the one with the highest negative average slope in the representation with the specific enthalpy in abscissa and the pressure in ordinates. This line marks the asymmetry in the specific heats at constant pressure, because for temperatures above this isotherm, the specific heat is higher in the high isóbara than in the low one, while below that isotherm the opposite occurs, and it is greater the specific heat in the low isobar. This isotherm that marks the asymmetry corresponds to the higher average value of the Vfp product along an isotherm between P0 and P1; within the limits marked by T0 and TM; and is located in the critical isotherm or slightly above, up to a few tens of degrees K.
When the temperature of the asymmetry coincides with the isotherm of the outlet temperature of the compressor, Tc, the worst situation in terms of performance occurs, as the hot spot, or the auxiliary, but at a temperature close to the output of the turbine , Tt, have to
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compensate for all the specific enthalpy that the regenerative exchanger cannot provide. And this is the enthalpy difference between Tc and T0 at pressure P0.
This proximity is limited by setting a maximum in the values of the logarithmic factor of isobar dilatation at the point of minimum enthalpy of the cycle; in addition to setting a minimum value for P0, above Pcr, to limit the non-recoverable enthalpy in the regenerative heat transfer. This enthalpy corresponds to:
AHnñ Vmaxfpmax Pl ^ o)
This value is notoriously dependent on P0, since fpmax can be identified with the maximum value of fp for the isobar in question. That is, the value acquired by fp in the cut between the isobar in question and the line of discontinuity, which was introduced as a novelty in this invention. This leads to establishing the requirement that P0 is equal to or greater than the pressure for which it is produced that the non-recoverable enthalpy is equal to the specific work of the turbine, that is
Vt0P0 „
tW, ~ (Pi-Po) = - (r ‘- i)
In turn Vmax and Vt0 are related by the equation
image19
AS ™ ax being the entropy, measured in P0, between the point of that isobar that has the asymmetry temperature, Ta, and the turbine outlet. The full prescription can be written as P0 is the pressure for which it is met
h
pmax
(rat - 1) gt (r - 1)
exp
ACE
t o
+ 1
Pt—)
p o
And as a general reference, with the usual values of the relevant fluids, the maximum value of fp is 20, at the cut-off point of P0 with the line of discontinuity. This sets the pressure value at the low isobar.
On the other hand, in the isentropic of compression lies a crucial part of the innovation, because a specific work of compression is sought that is low. At the same time you are looking for
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that, in absolute value, the specific enthalpy gained by the fluid when passing from point P1, Tc to P0, Tc, that is, along the isotherm Tc, is also moderate. This is not a thermodynamic line that represents a phase of the cycle, but a provision that is adopted to characterize it. Specifically, the prescription can be defined by the equality between both terms, which would be
VsÁP1 ~ Po) = VcfpÁPl ~ Po)
And since VSc and Vc correspond to two thermodynamic lines very close to each other, they will have very similar values, and therefore fpc must be 1, or close to this value, but not much higher.
As a consequence of this novel characterization of the thermodynamics of a non-condensing cycle, this invention has been configured, which is embodied in fundamental requirements to define the cycle. These are:
The value of the low isobar pressure, P0, which must be higher than the critical pressure, is chosen in such a way that the point at which it cuts to the T0 isotherm, which is the minimum temperature reached by the working fluid , have a value of the so-called logarithmic factor of isotary dilation less than 6; 1 being the reference value specified by this invention;
the value of the so-called logarithmic factor of isotary dilation, averaged along the isentropic compression, does not exceed 3; 0.5 being the reference value specified by this invention;
the value of the so-called isobar expansion logarithmic factor, averaged along the isothermal line of temperature equal to the turbine outlet temperature, from the low pressure P0 to the high pressure P1, does not exceed 2.5; 0.5 being the reference value specified by this invention;
the ratio between the high pressure P1 and low pressure P0, called the compression ratio, r, fulfills the closing ratio of the cycle
image20
■ ^ o
1 - gt - m
where gt is the exponent of the variation of the temperature, with respect to the pressure, in the expansion isentropic; m is the absolute value of the exponent of the variation of the specific volume, with respect to the pressure, in the isentropic compression; and A0 and A1 are respectively, for the low and high isobars, the value of the entropy increase
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multiplied by the average along the isobar, the quotient between the logarithmic factor of isobar dilatation, plus 1, and the specific heat at constant pressure.
The invention of the process is completed with the specifications of the components in which these cycles materialize, and particularly in the regenerative heat exchangers, which in the invention are not of a continuous piece, with two flows of the same fluid at different pressure in countercurrent, one that is heated, this being the high pressure, and another that is cooled, this being the low pressure, but in the invention, an auxiliary (non-regenerative) fluid heating flow is added to the regenerative exchanger working at high pressure, and the performance of this auxiliary flow is selected, always in countercurrent, so that it is continuously or discretely,
- and in the continuous mode the auxiliary heating flow is added as a third flow, which either brings heat to the low-pressure fluid, which on another contact surface heats the high-pressure fluid, or it provides heat directly to the high pressure fluid, by a contact surface different from that which separates the two flows of the same fluid;
- and in the discrete mode, the regenerative exchanger is divided into at least two parts, introducing between two consecutive parts an auxiliary countercurrent exchanger in which the low pressure fluid does not intervene, and the high pressure fluid is heated, starting of the auxiliary heating flow, which comes from an auxiliary source of lower temperature than the one that takes the working fluid to its maximum enthalpy state.
EXPLANATION OF THE FIGURES
The figures correspond in many cases to thermodynamic diagrams, either of a general type, or of a representative fluid, such as commercial refrigerant R-125, which is ethyl penta fluoride.
Figure 1 shows a diagram of an assembly of a device in which the process of the invention could materialize.
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Figure 2 shows a basic thermodynamic diagram, of R 125, using specific enthalpy as an abscissa and pressure, in logarithmic scale, as ordered.
Figure 3 shows a diagram of R 125 with exposure of contours of z and fp.
Figure 4 is similar to 3, but enlarging the high enthalpy part to reach the ideal gas domain.
Figure 5 includes in the diagram the trace of a cycle of the type prescribed in the invention.
Figure 6 shows only the phases of the cycle and their relevant elements, without the lines of the diagram.
Figure 7 shows two arrangements, (a) and (b) for mounting the third flow, or continuous auxiliary flow.
Figure 8 shows a regenerative exchanger sectioned in parts, with auxiliary (non-regenerative) heating carried out between consecutive sections.
To improve the understanding of the explanation of the figures, the elements that make up the invention are listed below:
1. Turbine (gas)
2. Turbine exhaust and connection to the low pressure circuit of the regenerative exchanger
3. Low pressure circuit of the regenerative exchanger
4. Regenerative exchanger
5. Low pressure circuit output of the regenerative exchanger
6. Working fluid circuit in the heat sink
7. Heat sink
8. Output of the working fluid from the heat sink, and connection to the compressor inlet
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9. Heat sink outside coolant inlet
10. Outlet of the external coolant from the heat sink
11. Compressor
12. Compressor outlet and connection to the high pressure circuit of the regenerative exchanger
13. High pressure circuit of the regenerative exchanger
14. Auxiliary caloric focus
15. Backflow circuit of auxiliary heating fluid
16. High pressure circuit output of the regenerative exchanger and hot spot connection
17. Hot spot
18. Output of the fluid at high pressure and at the highest temperature, for turbine entry
19. In the straight section of the regenerative heat exchanger with continuous auxiliary heating from the auxiliary to the low pressure flow, a tube through which the fluid flows at high pressure.
20. In the straight section of the regenerative heat exchanger with continuous auxiliary heating from the auxiliary to the low pressure flow, annular section conduction through which the fluid circulates at low pressure.
21. In the straight section of the regenerative heat exchanger with continuous auxiliary heating from the auxiliary to the low pressure flow, annular section conduction through which the auxiliary fluid circulates.
22. In the variant given for numerical labels 19, 20 and 21, heat flow from the low pressure to the high fluid.
23. In the variant said in 22, heat flow from the auxiliary to the low pressure fluid.
24. Applicable to the variant in which the auxiliary flow directly heats the fluid at high pressure, and is the mark that indicates that heat flow. In this case, represented in Figure 7b, the high pressure fluid circulates through the annular conduit 19b, and the low pressure fluid, through the central tube 20b, in this case the heat passing, as indicated in 22b, from the tube central to the ring around it.
25. Main hot spot
26. Auxiliary hot spot
27. High pressure fluid auxiliary heaters, according to LT1 line
28. Sections or parts of the regenerative exchanger.
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In the diagrams there are several lines defined by a property, or because the value of a given variable is constant. The letter H always represents specific enthalpy in kJ / kg, be it uppercase or lowercase. The labels used in the figures are:
LB: line under which the liquid and vapor phases coexist, that is, biphasic zone.
LD: secondary discontinuity line, in which the highest value of the f factor for each pressure is reached.
Pcr: critical point.
To make the enthalpy-pressure diagrams of Figures 2,3 and 4 unequivocal, five sets of labels are used, corresponding to a letter followed by a number, the letters being
V: specific volume T: temperature S: entropy
Z: compressibility factor F: logarithmic factor of isobar dilatation The complete specification of the labels shown is:
V1 = 0.0008 m3 / kg V2 = 0.001 m3 / kg V3 = 0.002 m3 / kg V4 = 0.005 m3 / kg V5 = 0.015 m3 / kg V6 = 0.04 m3 / kg T1 = 0 ° C T2 = 100 ° C T3 = 200 ° C T4 = 300 ° C T5 = 400 ° C T6 = 500 ° C
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T7 = 10 ° C T8 = 32 ° C T9 = 66 ° C T10 = 99 ° C T11 = 134 ° C S1 = 1.2 kJ / kgK S2 = 1.6 kJ / kgK S3 = 2 kJ / kgK S4 = 2 , 4 kJ / kgK S5 = 1 kJ / kgK S6 = 1.2 kJ / kgK S7 = 1.4 kJ / kgK S8 = 1.6 kJ / kgK S9 = 1.8 kJ / kgK Z1 = 0.1 Z2 = 0.2 Z3 = 0.4 Z4 = 0.6 Z5 = 0.8 Z6 = 1 F1 = 19 F2 = 7 F3 = 3 F4 = 1 F5 = 0
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Also in the figures in which the cycle is represented, or part of it, such as partially regenerative heating, labels with the following acronyms are used:
LP for isobars. The LP0 mark indicates precisely the line in which the pressure is constant and is worth P0. LP1 is the high isobar. They are also used as labels to designate said isobars P0 and P1.
LT for isotherms, as is the case with the highest inclination in the diagram (enthalpy, log P) which is the LTa line. Note that the LTpa label is a line parallel to the LTa line. The labels Tpa and Ta are also used to designate these last two isotherms respectively; and also Tt to designate the isotherm of the turbine exhaust temperature, and Tc for that of the drive from the compressor.
T0 is the temperature of the minimum enthalpy point of the cycle, and TM the maximum enthalpy point.
LSc is the isentropic compression (also as Sc).
LSt is the isentropic turbine expansion (also denoted St)
In Figure 8, LTP0 is the cooling line, up to Ta, of the low isobar; and LTP1 is the heating line of the high isobar, which combines regenerative parts, receiving heat from the lower isobar, with parts in which it is heated with the auxiliary fluid.
EMBODIMENT OF THE INVENTION
The invention is embodied by having a set of thermal engineering elements or components, ranging from the compressor (11) to the turbine (1), through the regenerative exchanger (4), plus the means (14, 17) for heating the high pressure fluid with the auxiliary means, either continuously, or jumping. The materialization includes as an essential matter to fix the levels of the relevant variables in the definition of the cycle, complying with the prescriptions already expressed. These levels depend substantially on the working fluid used.
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Figure 1 shows a scheme that visualizes a preferred embodiment of the invention. The device on which the thermodynamic cycle is performed 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. The outlet (5) of said low pressure circuit (3) is connected to a heat sink (7) through the corresponding circuit (6) of the working fluid in said sump. This heat sink (7) is refrigerated, with an inlet (9) and an outlet (10) of the external refrigerant of the heat sink. The outlet (8) of the working fluid from the heat sink (7), sets the conditions under which said working fluid enters the compressor (11).
Once the working fluid has finished its evolution in the compressor (11), the outlet (12) of the compressor (11) assumes the connection to the high pressure circuit (13) of the regenerative exchanger (4). Said regenerative exchanger (4) has an auxiliary heat source (14) that feeds the exchanger (4) by means of a countercurrent circuit (15). Thus, when the working fluid evolves through the outlet (16) of the high pressure circuit (13) of the regenerative exchanger (4), it is able to begin its evolution in the hot spot (17), increasing its enthalpy. At the outlet (18) of said hot spot (17), the fluid is in a position to evolve in the turbine (1), thus initiating a new cycle.
The main characteristics of the characteristic thermodynamic cycle of this invention are shown in Figure 6.
Different embodiments of the regenerative exchanger (4) acting in continuous mode are shown in Figures 7a and 7b. Figure 7a shows different sections through which fluids circulate; thus, reference (19) indicates the high pressure fluid circulating through the exchanger (4); reference (20) indicates the low pressure fluid circulating through the exchanger (4), and reference (21) shows the section through which the auxiliary fluid from the heat source (14) circulates. The reference (23) represents the heat transfer that occurs from the auxiliary fluid circulating through the section (21) to the low pressure fluid circulating through the section (20). Similarly, the reference (22)
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represents the heat transfer that occurs from the low pressure fluid that circulates through the section (20) to the high pressure fluid that circulates through the section (19).
Figure 7b shows another configuration on the operation of the regenerative exchanger (4) in the continuous mode of operation. Reference (19b) indicates the high pressure fluid that circulates through the exchanger (4); reference (20b) indicates the low pressure fluid circulating through the exchanger (4), and reference (21) shows the section through which the auxiliary fluid from the heat source (14) circulates. The reference (24) represents the heat transfer that occurs from the auxiliary fluid circulating through the section (21) to the high pressure fluid circulating through the section (19b). Similarly, the reference (22b) represents the heat transfer that occurs from the low pressure fluid that circulates through the section (20b) to the high pressure fluid that circulates through the section (19b). That is, in Figure 7b the high pressure fluid confined in section (19b) is "surrounded internally" by the low pressure fluid of section (20b), and "externally" by the auxiliary fluid of section ( 21) from the heat source (14). In both the case of Figure 7a and that of Figure 7b, what is achieved is to provide the high-pressure fluid with the complementary heat that the regenerative heating cannot transfer. This heat is supplied by an auxiliary source, at a lower temperature than the main source.
Figure 8 shows the embodiment of the regenerative exchanger (4) acting in discrete mode. The regenerative heat exchanger is divided into sections (28), introducing between them an auxiliary heat exchanger (27) in which the high pressure fluid is heated from an auxiliary flow from an auxiliary source (26 ) of lower temperature than the main temperature (25).
The materials used to construct these elements will generally be of conventional metals, since the working fluid does not have to accompany very intense properties in corrosion and the like, but rather on the contrary, manifest quite inert properties; Examples of this are CO2 and R-125 refrigerant (which is ethyl pentafluoride).
The basic requirement to be met by the working fluid is that its critical temperature is several degrees Celsius higher than the highest ambient temperature to cool. In that the two mentioned work fluids have notable differences, because the
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The critical temperature of CO2 is 31 ° C and that of R-125 is 66 ° C. There is therefore much more room in general to work with the latter, which also has a critical pressure less than half of CO2, which is almost 74 bar (7.37 MPa), while that of R-125 is 36 Pub.
As an example, if a cycle is taken in which R-125 is used as a working fluid, defined by a minimum temperature T0 = 50 ° C, a maximum temperature TM = 500 ° C, a low pressure slightly above Critically, P0 = 1.01Pcr, and a compression ratio r = 3, it is observed that the maximum efficiency that can be obtained without adding auxiliary sources of heat is 0.41.
However, considering that the cost of the heat source increases proportionally to the square of the temperature (which is easily imaginable if one thinks of solar thermal applications where increasing the temperature of the source considerably increases the cost of the equipment necessary for this purpose) , it would be obtained that, for the mentioned cycle, adding an auxiliary source (either for continuous or discrete use) at a temperature of 250 ° C, would result in an efficiency equal to 0.57, and in the case of adding two auxiliary sources, one at 200 ° C and another at 325 ° C, the efficiency would be 0.61, that is, almost 50% higher than that achieved without adding the auxiliary sources described in the invention.
In the previous example, in order to consider the cost of temperature, the performance has been defined as follows
and
Wt-Wc
ZQ-T2
1 m
2
Wt being the specific work done by the turbine, Wc the specific work consumed by the compressor, Q the heat contributed by each source, at its temperature T, which only equals TM in the case of the main heating source.
Although CO2 offers less temperature range and requires higher pressures, it should not be discarded. Keep in mind that you can get closer T0 to Tcr if you raise a little P0 over Pcr, since the temperature at which the maximum enthalpy variation occurs is increased.
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If the isobar of 8 MPa is considered, the fp values for the temperatures of the range of interest are
 T (° C)  fp
 17.5  1.7
 22.5  2.7
 27.5  5.7
Even at the limit of T0 = 30 ° C, which would be on the edge of the fundamental prescription, it is found that in its isentropic from 8 MPa it reaches 20 MPa (that is, with r = 2.5) with a temperature of 50 ° C, which represents a thermal exponent gc = 0.07 that is acceptable (but also to the limit). Also, the density rises from 700 kg / m3 to 775, which means that the value of the exponent m = 0.111.
Precisely because it is so on the limit, the specific enthalpy difference between the point of T = 50 ° C and T = 30 ° C at 8 MPa is very high, 135 kJ / kg. This heat must be extracted precisely with the cold outside focus, and must be replenished with the auxiliary fluid in the high-pressure fluid, before it receives the final heating in the main focus.
If a TM = 773 K is considered, with a CO2 behavior in the ideal gas turbine, it would leave (expanding from 20 MPa to 8) to 630 K; and produce a specific job of 120 kJ / kg. The compression work would be 15 kJ / kg. Without considering any type of thermal losses, the yield would be (120-15) / (120 + 135) = 0.41.
The above data show the breadth of possibilities that exist to materialize this invention, which is mainly based on preparing a map of the values of the fp factor in the design window of interest, depending on the availability of heat source, specified in TM and outside cooling, in T0.
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It is now up to specify the auxiliary heating fluid prescriptions, which are specified above all in the mass flow or expense (in kg / s) that must flow to make the regenerative complementary heating possible. In the case of the continuous mode of operation, the regenerative heat exchanger (4) has an auxiliary heat source (14) that feeds said exchanger (4) by means of a countercurrent circuit (15), and is called m'x at Mass expenditure of the auxiliary flow, whose temperature is denoted by Tx and has a specific heat at constant pressure Cpx, and m 'to the mass expenditure of the working fluid of Brayton, and Cp1 and Cp0 at specific heats at constant pressure in the isobars of high and low, the expense m'x is
mx = m
Cpi Cp0
-‘Px
In the case of the device shown in Figure 7 (a), in the regenerative heat exchanger (4):
or there is a heat transfer (23) from an auxiliary fluid from a focus (14), which circulates through a section (21) of the exchanger (4), to the low pressure fluid (3) that circulates through another section ( twenty);
or a heat transfer (22) occurs from the fluid of the low pressure circuit (3) that circulates through said section (20), to the fluid of the high pressure circuit (13) that circulates through another section (19) of the exchanger (4)
and the thermal hierarchy Tx> T0> T1 is given between the temperatures of the auxiliary flow, Tx, the low pressure flow T0 and the high pressure T1, the relationship between them depending on the values of the Number of Transmission Units, N, of each heat transmission interface, between the auxiliary flow and the low pressure flow, Nx0 and between it and the high flow N01, each N in the numerator having the overall heat transmission coefficient multiplied by the transmission area, and the denominator it is the smallest of the products, of the two fluids involved, of the mass expenditure for specific heat at constant pressure, and the relationship is fulfilled
Tn - 7, =
N.
xO
Nx0 + N,

01
which prescribes the required value of Tx as the numbers of transmission units in the interfaces are sized, the three temperature lines being parallel to each other.
For the device of Figure 7 (b), the high pressure flow with temperature T1 is heated
on two independent faces, one from the low pressure flow with T0 and one from the
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auxiliary with Tx; and the exchanger as a whole works in a balanced way, with parallel evolutions in the three temperatures. In the regenerative heat exchanger
(4):
or there is a heat transfer (24) from an auxiliary fluid from a focus (14), which circulates through a section (21) of the exchanger (4), to the high pressure fluid (13) that circulates through another section ( 19b);
or there is a heat transfer (22b) from the low pressure circuit fluid (3) that circulates through a section (20b) of the exchanger (4), to the high pressure circuit fluid (13) that circulates through another section (19b) of the exchanger (4),
and all three temperatures meet
T0-Ti = V0-Tx)
N.
N,
Xl
Xl
~ N,
01
Nx1 being the number of transmission units between the auxiliary fluid, whose temperature is Tx, and the high pressure fluid, of temperature T1, and N01 being the number of transmission units between the low pressure fluid, whose temperature is T0, and high pressure fluid; and when Nxl = N01, the solution is T0 = Tx, or vice versa.
Once the invention is clearly described, it is noted that the particular embodiments described above are subject to modifications in detail as long as they do not alter the fundamental principle and essence of the invention.

权利要求:
Claims (9)
[1]
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1. Cyclic thermodynamic process without condensation of the fluid and with prescriptions bounded on its points of minimum and maximum enthalpy, where the cycle works between an isobar of lower pressure, or low isobar, which is at P0, and a high isobar, or higher pressure, P1, existing:
- a compression phase, in which a compressor (11) aspirates the fluid at its point of lowest specific enthalpy of the entire cycle, at pressure P0 and temperature T0, and elevates it from pressure throughout an isentropic evolution, up to P1, leaving the compressor (11) with a temperature Tc;
- following a heating phase;
characterized in that the pressure P0 is higher than the critical pressure of the working fluid, and the temperature T0 is lower than the critical temperature of said fluid, and because in said heating phase three simultaneous types of heat sources concur, which are:
- the fluid itself, in another phase of the cycle, in which it is hotter than the fluid itself at its exit from the compressor (11),
- a hot spot (17), the main source of heat contribution to the working fluid, with which the fluid is heated to the maximum temperature reached by the TM working fluid,
- an auxiliary thermal bulb (14), which preheats the working fluid without reaching the TM temperature, in any case;
existing in addition:
- 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 to the pressure P0, leaving the turbine (1) or expanding machine where perform this phase with a temperature Tt;
- following a cooling phase in which two types of cooling actions concur, which are:
or the fluid itself, in another phase of the cycle in which it is colder than the fluid itself at its exit from the turbine (1),
or the external cooling sump, which cools the fluid to T0.
[2]
2. Thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, as set forth in claim 1,
characterized in that the prescriptions are:
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the low isobar pressure, which is the pressure of the minimum enthalpy point, P0, has a maximum value of the so-called logarithmic factor of isotary dilation less than 20, giving this value in the cut of the isobar with the secondary discontinuity line that present fluids above the critical point; the ratio between the maximum and minimum fluid temperatures, TM / T0, must be greater than the compression ratio r, corresponding to the P1 / P0 ratio, raised to the sum of the thermal exponents in the isentropic compression and expansion evolutions; The value of the low isobar pressure, P0, which must be higher than the critical pressure, is chosen in such a way that the point at which it cuts to the T0 isotherm, which is the minimum temperature reached by the working fluid , have a value of the so-called logarithmic factor of isotary dilation less than 6; 1 being the reference value specified by this invention;
the value of the so-called logarithmic factor of isotary dilation, averaged along the isentropic compression, does not exceed 3; 0.5 being the reference value specified by this invention;
the value of the so-called isobar expansion logarithmic factor, averaged along the isothermal line of temperature equal to the turbine outlet temperature, from the low pressure P0 to the high pressure P1, does not exceed 2.5; 0.5 being the reference value specified by this invention;
and the high pressure P1 is prescribed because the ratio between said high pressure P1 and the low pressure P0, called the compression ratio, r, meets the cycle closing ratio
image 1
■ ^ o
1 - gt - m
where gt is the exponent of the variation of the temperature, with respect to the pressure, in the expansion isentropic; m is the absolute value of the exponent of the variation of the specific volume, with respect to the pressure, in the isentropic compression; and A0 and A1 are respectively, for the low and high isobars, the value of the entropy increase multiplied by the average along the isobar, of the ratio between the logarithmic factor of isobar expansion, plus 1, and the specific heat at pressure constant.
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[3]
3. Cyclic thermodynamic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to claim one or two, characterized in that the heating phases taking place with:
or the working fluid itself, in another phase of the cycle, in which it is hotter, or an auxiliary thermal bulb (14), which preheats the working fluid but without reaching the TM temperature, in any case,
they are carried out in a regenerative exchanger (4) to which an auxiliary, non-regenerative, heating flow of the high-pressure working fluid is added, this auxiliary flow always acting in countercurrent (15) with respect to the circuit (13) of the flow fluid High pressure work.
[4]
4. Thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to claim 3, characterized in that the regenerative exchanger (4) works in a continuous mode, adding a third heating flow, coming from an auxiliary source (14), which either provides heat (23) to the low pressure fluid, which by another contact surface provides heat (22) to the high pressure fluid; or it provides heat (22b, 24) directly to the high pressure fluid, by a different contact surface from the one that separates the two flows of the same fluid.
[5]
5. Cyclic thermodynamic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points according to claim three, characterized in that the regenerative exchanger (4) works in a discrete mode, said regenerative exchanger (4) being sectioned in, at less, two parts, introducing between said two consecutive parts a countercurrent auxiliary exchanger in which the low pressure fluid does not intervene, and the high pressure fluid is heated, from another auxiliary fluid that receives heat from an auxiliary source (14) of lower temperature than the one that takes the working fluid to its maximum enthalpy state.
[6]
6. Device for carrying out a thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to any of the first to fifth claims, characterized in that said device on which the thermodynamic cycle is performed understands:
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- a regenerative heat exchanger (4), which integrates at least one high pressure circuit (13) and a low pressure circuit (3);
- a gas turbine (1), whose exhaust (2) is connected to the low pressure circuit inlet (3) of the regenerative exchanger (4);
- 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) external refrigerant;
- 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 (13) of the heat exchanger regenerative heat (4);
- a hot spot (17), whose input is connected to the output (16) of the high pressure circuit (13) of the regenerative heat exchanger (4), and whose output (18) is connected to the gas turbine (1) ).
[7]
7. Device for carrying out a thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to claim 6, characterized in that the regenerative heat exchanger (4) has a heat source auxiliary (14) that feeds said exchanger (4) by means of a countercurrent circuit (15), and is called m'x to the mass expenditure of the auxiliary flow, whose temperature is denoted by Tx and has a specific heat at constant pressure Cpx, and m 'to the mass expenditure of the working fluid of the Brayton, and Cp1 and Cp0 to the specific heats at constant pressure in the high and low isobars, the expenditure m'x is
mv = m
-p i
-p o
px
[8]
8. Device for performing a thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to any of the sixth or seventh claims, characterized in that, in the regenerative heat exchanger (4) :
or there is a heat transfer (23) from an auxiliary fluid from a bulb (14), which circulates through a section (21) of the exchanger (4), to the low fluid
pressure (3) circulating through another section (20);
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or a heat transfer (22) occurs from the fluid of the low pressure circuit (3) that circulates through said section (20), to the fluid of the high pressure circuit (13) that circulates through another section (19) of the exchanger (4)
and the thermal hierarchy Tx> T0> T1 is given between the temperatures of the auxiliary flow, Tx, the low pressure flow T0 and the high pressure T1, the relationship between them depending on the values of the Number of Transmission Units, N, of each heat transmission interface, between the auxiliary flow and the low pressure flow, Nx0 and between it and the high flow N01, each N in the numerator having the overall heat transmission coefficient multiplied by the transmission area, and the denominator it is the smallest of the products, of the two fluids involved, of the mass expenditure for specific heat at constant pressure, and the relationship is fulfilled
Tt =
N,
xO
Nx0 + N,
■ Vx-YOU
01
image2
which prescribes the required value of Tx as the numbers of transmission units in the interfaces are sized, the three temperature lines being parallel to each other.
[9]
9. Device for carrying out a thermodynamic cyclic process without condensation of the fluid and with prescriptions bounded on its minimum and maximum enthalpy points, according to any of the sixth or seventh and eighth claims, characterized in that, in the regenerative heat exchanger (4):
or there is a heat transfer (24) from an auxiliary fluid from a focus (14), which circulates through a section (21) of the exchanger (4), to the high pressure fluid (13) that circulates through another section ( 19b);
or there is a heat transfer (22b) from the low pressure circuit fluid (3) that circulates through a section (20b) of the exchanger (4), to the high pressure circuit fluid (13) that circulates through another section (19b) of the exchanger (4),
and the exchanger as a whole works in a balanced way, with parallel evolutions in the three temperatures, which meet
To-T ^ (T0 - Tx)
Nxi
Nxi ~ N0Í
Nx1 being the number of transmission units between the auxiliary fluid, whose temperature is Tx, and the high pressure fluid, of temperature T1, and N01 being the number of transmission units between the low pressure fluid, whose temperature is T0, and high fluid
Pressure; and when Nxl = N01, the solution is T0 = Tx, or vice versa.
32

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同族专利:

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公开号 | 公开日

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ES2652522B2|2019-01-16|

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US20160010551A1|2014-07-08|2016-01-14|8 Rivers Capital, Llc|Method and system for power production wtih improved efficiency|ES2776024A1|2020-03-03|2020-07-28|Univ Madrid Politecnica|THERMODYNAMIC SYSTEM WITH CLOSED CYCLE, WITH REGENERATIVE REFRIGERATIONS TO COUNTER CURRENT, TO GENERATE MECHANICAL ENERGY IN ONE OR MULTIPLE AXES, FROM EXTERNAL FLOWS OF HOT FLUIDS |

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