JPH08278090A - Heat exchanger and heat machine - Google Patents

Abstract

PURPOSE: To reduce the thermal conduction in the flowing direction of fluid, to make the fluid pressure in a pore group or groove group uniform and to increase the heat exchange capacity by opening the pore group or the groove group for reducing the thermal resistance between the fluid and a thin plate on the plate, providing a small gap in the flowing direction of the fluid and providing a plurality of the gaps in parallel on the plate. CONSTITUTION: A plurality of thin plates 30 made of metal material having high thermal conductivity which is spread perpendicular to the flowing directions of fluid 1 (F1) and fluid 2 (F2) are provided substantially in contact with pressure resistant circular tubes 10 and 20. A plurality of pores (circular pores) 32 for passing the fluids 1, 2 are opened to a pore group on the plate 30 to reduce the thermal resistance between the fluid and the plate 30. The heat from the fluid 1 is rapidly transferred to the tube 10 by the highly thermal conduction in the plate 10. As a result, the heat exchange between the fluids 1 and 2 is increased as compared with the case that the plate 30 is not provided. A laminated structure is formed via a small gap 34 to make the fluid pressure in the hole 32 uniform.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a heat exchanger and a heat machine, and more particularly to a heat exchanger having a very large heat exchange capacity per unit volume of a device and a heat machine using the same.

[0002]

2. Description of the Related Art As a conventional heat exchanger, for example, a multi-tube heat exchanger widely used as an important element of a heat engine or a heat pump is known.

[0003] In a multi-tube heat exchanger, a plurality of straight circular tubes are arranged in parallel in one direction, a first fluid flows in the inside of each circular tube in the axial direction of the circular tube, and outside the circular tube. The second fluid is caused to flow in the axial direction of the circular tube so that heat exchange (transfer of heat) is performed between the two fluids. In a heat engine and a heat pump, a heat exchanger is used in the overall apparatus. Often they occupy a large volume fraction and the fluid flow is often turbulent.

[0004]

There have been many experimental and theoretical studies on the flow of fluid and heat in the conventional multi-tube heat exchanger described above. ,
For example, as can be seen in the Handbook of Mechanical Engineering (published by the Japan Society of Mechanical Engineers), in these conventional studies and the conventional design guidelines for the multitubular heat exchanger based on these studies, the following points of interest of the present invention, namely, The theoretical fact that the volume of a heat exchanger can be significantly reduced by reducing the diameter (radius) of a circular pipe (or a circular hole described later) through which a fluid flows has not been recognized as practical, and thus this Utilizing the fact, no heat exchanger having a very large heat exchange capacity per unit volume of the device was designed and manufactured, and the options for heat exchanger design were narrowed.

The present invention has been made in view of the above-mentioned problems of the prior art, and an object of the present invention is to make use of theoretical facts which have not been noticed so far to achieve the conventional art. An object of the present invention is to provide a heat exchanger and a heat machine having a very large heat exchange capacity per unit volume, which has been regarded as difficult.

[0006]

In order to achieve the above object, a heat exchanger according to the present invention exchanges heat with each other.
The two fluids are separated by a pressure-resistant wall made of a material that can withstand the pressure difference between the two fluids, and the two fluids flow in opposite directions to each other, and are separated by at least one of the two fluids. A thin plate made of a material having a high thermal conductivity that is substantially in contact with the pressure-resistant wall and spreads in a direction perpendicular to the direction in which the fluid flows, and the thermal resistance between the fluid and the thin plate is increased by passing the fluid through the thin plate. Drilling a group of pores or grooves to reduce
The thin plates were arranged in parallel with a small gap in the flow direction of the fluid to reduce the heat conduction in the flow direction of the fluid and to equalize the fluid pressure in the group of pores or grooves. It is characterized by the following.

In the heat exchanger according to the present invention, the pressure-resistant wall is at least one pressure-resistant tube through which one of the two fluids flows. It is characterized by being constituted by the tube wall of the pressure-resistant tube arranged inside the pressure-resistant tube through which the other fluid flows.

Further, in the heat exchanger according to the present invention, in the heat exchanger according to the second aspect, each of the pressure-resistant walls includes a member having a closed curved cross section and a height of about the small gap. It is characterized by being arranged between thin plates.

Further, in the heat exchanger according to the present invention, at least a part of the pores or the grooves of the adjacent thin plates are arranged at positions shifted from each other. It is characterized by comprising.

[0010] Further, the thermal machine of the present invention makes the working gas thermally contact between the heat source and the two heat sources having a substantially constant temperature.
A thermal machine that approximately realizes an Ericsson cycle without loss of entropy generation by performing isothermal compression, isothermal expansion, and isobaric heat exchange on the working gas,
A heat exchanger according to any one of claims 1 to 4 is provided.

[0011]

In the heat exchanger according to the present invention, one of the fluids (fluid 1) exchanges heat with the thin plate when passing through a group of pores or grooves formed in the thin plate. The exchanged heat is transmitted to the pressure-resistant wall by heat conduction in the thin plate material having a high thermal conductivity, and is directly transmitted to the other fluid (fluid 2) flowing in contact with the pressure-resistant wall, or similarly to the case of the fluid 1, Is transmitted to the group of pores or grooves formed in the thin plate by heat conduction in the thin plate material which spreads in the direction perpendicular to the axis of the pressure tube toward the fluid 2 while being in contact with the pressure tube. Alternatively, the flow is transmitted to the fluid 2 passing through the groove group, but is described as a laminar flow (hagen-poiseuille flow) in a circular hole described later.
According to the heat conduction theory, the volume V of the heat exchanger is equal to the radius R of the hole.
₀ or the groove width 2X ₀ can be reduced in proportion to R ₀ ² by selecting a small value, so that it is possible to take into account the difficulty of drilling holes or grooves, cost, clogging of holes or grooves due to dust, etc. By adopting the smallest possible hole radius R ₀ or groove width 2X ₀ , a heat exchanger having an extremely large heat exchange capacity per volume can be obtained.

By employing such a group of pores or grooves, it is possible to realize a heat exchanger which is easier to manufacture and has a significantly smaller volume than a multi-tube heat exchanger using circular tubes. In addition, by having a plurality of thin plates having a group of pores or grooves arranged side by side with a small gap, heat conduction in the flow direction of the fluid is reduced, and other than the hydrostatic stress applied to the thin plate. The stress can be reduced to extend the life of the thin plate.

[0013]

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is an exploded perspective view of a heat exchanger and a heat machine according to an embodiment of the present invention. An example of the structure will be described.

FIGS. 9 and 10 show the main parts of an example 50 of such a conventional multi-tube heat exchanger.

The fluid 1 is caused to flow in the z direction in the internal space 11 of one pressure-resistant circular tube 10, and the pressure-resistant circular tube 10 is placed inside the one pressure-resistant circular tube 20 concentrically with respect to the pressure-resistant circular tube 20. It is placed outside the pressure-resistant tube 10 and at the same time as the pressure-resistant tube 2
Heat is exchanged between the fluid 1 and the fluid 2 by flowing the fluid 2 in the −z direction into the space 13 inside the zero.

FIGS. 11 and 12 show another example of the arrangement of pressure-resistant tubes in an example 50 of a conventional multi-tube heat exchanger, in which seven pressure-resistant tubes 10 are replaced with one pressure-resistant tube. 20, the fluid 1 flows in the z direction in the interior 11 of each pressure-resistant tube 10, and the fluid 2 flows into the space 13 outside the pressure-resistant tube 10 and inside the pressure-resistant tube 20. Heat is exchanged between the fluid 1 and the fluid 2 by flowing in the −z direction.

FIGS. 1 and 2 show a main part of a heat exchanger 100 according to a first embodiment of the present invention. One or both of the spaces 13 inside the tube 20 (both in FIGS. 1 and 2) substantially contact the pressure-resistant circular tube 10 and the pressure-resistant circular tube 20, and the fluid 1 (F1) and the fluid 2 (F2)
A plurality of thin plates 30 made of a metal material having a high thermal conductivity (see Table 1) spreading in a direction perpendicular to the flowing direction (+ -z direction) are placed on the thin plate 30. By forming holes (circular holes) 32 to form a group of pores, the thermal resistance between the fluid and the thin plate 30 is reduced, and the heat from the fluid 1 is rapidly released from the fluid 1 by a high degree of heat conduction in the thin plate 30. 1
0, and as a result, the heat exchange between the fluid 1 and the fluid 2 is greatly increased as compared with the case where the thin plate 30 is not provided.
In a structure in which the small holes 34 are stacked in a small gap (−z direction), the gap reduces heat loss due to heat conduction in the direction in which the fluids 1 and 2 flow (+ −z direction), and the circular holes 32. Equalize the fluid pressure in the
The occurrence of stress due to non-uniform pressure applied to zero is reduced.

In the heat exchanger 100, as shown in FIG. 5, the thin plate 30 spreading over the width of the fluid 1 is continuously spread in the fluid 2 as it is, and a plurality of thin pressure-resistant tubes 15 are used instead of the pressure-resistant circular tube 10. be able to. In this case, the thin plates 30 and the annular thin pressure-resistant tubes 15 are alternately stacked, and the thin plates 3
An annular groove 31 is provided at 0 and the pressure-resistant tube 15 is fitted therein, and the fitted portion is welded or brazed to prevent leakage. This structure further ensures the heat transfer from the fluid 1 to the fluid 2.

FIGS. 3 and 4 show another example of the main part of the heat exchanger 100 according to the present invention.
In one or both of the internal space 11 and the space 13 outside the pressure-resistant circular tube 10 and inside the pressure-resistant circular tube 20, pores (circles) similar to those shown in FIGS. (Hole) 3
A plurality of thin plates 30 with two groups are provided, and a plurality of pressure-resistant circular tubes 10 are provided. Considering that the number N ₁ of the pressure-resistant pipes 10 is to keep the entropy generation loss in the radial direction within the inside 11 of the pressure-resistant pipe 10 within a certain limit, the following equation (4) is used.
Preferably, it is determined according to 6).

Also, in the embodiment shown in FIGS. 3 and 4, the thin plate 30 spreading in the fluid 1 is continuously spread in the fluid 2 as shown in FIG. As in the case described above, a thin annular pressure-resistant tube 15 can be used.

The outline of the structure of the heat exchanger according to the present invention has been described above. Here, the theory of heat conduction in the Hagen-Poiseuille flow, which is the theoretical basis of the idea of the present invention, in particular, pores (circular holes) 32 choosing the size of (the radius R ₀ of the circular holes) reduce the volume V of the heat exchanger will be described the theoretical basis for being able to be reduced in proportion to the R ₀ ² by the.

First, when the flow in the circular hole 32 is represented by a velocity vector (u, v, w) on a rectangular coordinate system (x, y, z) with the circular axis as the z-axis, (u , v, w) = (0,0 , w (r)) r 2 = x 2 + y 2 (1) w (r) = 2W A (1- (r / R 0) 2) (2) W A = −P _z R ₀ ² / (8η) = − ZP _z L / (8η) (3) L = R ₀ ² / Z (4) P _z = (∂P /） _z ) = gradP = constant (5) Assuming that this holds, this flow vector is the Navier-Stokes equation governing the flow of Newtonian fluid D (u, v, w) / Dt = -gradP + ηΔ (u, v, w) D / Dt = (∂ / ∂ t) + u (∂ / ∂x) + v (∂ / ∂y) + w (∂ / ∂z) (6) Alternatively, the expression (6) can be expressed in accordance with the flow only in the axial direction with respect to the axial object in the circular hole. Rewritten P _z / η = (1 / r) (∂ / ∂r) (r (∂W / ∂r)) (7) and boundary condition W / ∂r = 0 when r = ₀ and W when r = R ₀ = 0 (8) is satisfied. This is known as the Hagen-Poazoille flow. Here, t is time, r is a distance variable in the radial direction of the circular hole 32, η is the viscosity coefficient of the fluid, and R ₀ is the circular hole 32.
Radius, Z is intended to be the axial length of the circular hole 32 to represent the total sum of the thickness of the thin plate 30, W _A is the average flow velocity of the circular hole in the 32, L is characterized lengths of formula (4) It is assumed that P is a static pressure, and _Pz is a derivative of P in the z direction and is a constant.

Further, when T, ρ, C _p , and κ are temperature, density, specific heat at constant pressure, and heat conductivity of the fluid, respectively, the temperature distribution in the circular hole 32 is represented by a heat conduction equation ρC _p (DT / Dt) = κΔT + (u, v, w) ∇P (9) Table 2 below shows the physical properties of typical working fluids.

[0024]

[Table 1]

[0025]

[Table 2] More specifically, the second term (the viscous heat generation term) on the right side of the above equation (9) is omitted, and this equation is rewritten into an expression conforming to the shape of the circular hole 32, and T is defined as the temperature difference from the hole wall. _{T R T (r, z)} = T z (z) + T R (r) (10) to replace equation _{wT z / a = (1 /} r) (∂ / ∂r) (r (∂T R / ∂ r)) (11) and boundary conditions The temperature distribution in the circular hole 32 follows ∂T _R / ∂r = 0 at r = ₀ and T _R = 0 at r = R 0 (12). Now the temperature distribution _{T z = (∂T z (z} ) / ∂z) = const. (13) T R = (18/11) T RA (1- (4/3) (r / R 0) 2 + ( _{1/3) (r / R 0)} 4) (14) T RA = 11P z T z R 0 4 / (384 aη) (15) a = k / (ρC p) (16) = k (1-1 / Γ) T _A / P _A (in the case of a perfect gas) (16 ′) satisfies the equation (11) and the equation (12), and is accompanied by the viscosity of the fluid in the second term on the right side of the differential equation (9) of heat conduction. Satisfies the case where heat generation is ignored. Here, T _z (Z) is the temperature of the inner wall of the circular hole 32, T _RA is the average temperature difference obtained by weighting the temperature difference between the fluid and the hole wall by the flow velocity w, and T _A and P _A are inside the heat exchanger. Mean temperature and pressure of the fluid. Since the second term on the right side of the equation (9) is smaller than the other terms, the temperature distribution of the equations (12) to (16 ′) is a sufficiently satisfactory approximate solution of the equation (9). The viscous loss term of (9) (the second term on the right side) can be calculated separately.

For the Hagen-Poiseuille flow of equations (2) and (3), this viscous loss term is (u, v, w) ∇P = | W _A P _z | = 8ηW _A ² R ₀ ^-2 = P _z ² R ₀ ² / (8η) (17), and when a value obtained by integrating the viscous heat loss with respect to the entire volume of the heat exchanger is E, F = G / ρΕ = FZP _z (18). However, G is the mass flow rate of the fluid, F is a volumetric flow rate of the fluid under average temperature and pressure conditions in the heat exchanger (T _A, P _A). Here, as described above, this viscous heat loss is discussed as being sufficiently small and negligible, but it will be later confirmed numerically that it is actually negligible.

The above analysis is known [Sellers, JR et al., Trans. ASME, 78-2, 441 (195)
6)]. The feature length L for a product ^{_{Lw A L = R 0 2 /}} Z = (384 aηT RA) 1/2 (11ZT z ZP z) -1/2 (19) of the formula (4) R ₀ = (ZL ^{) 1/2 = Z 1/2 (384 aηT} RA) 1/4 (11ZT Z ZP z) -1/4 (20) Lw A = 48aT RA / (11ZT z) (21) holds. Here, ZT _z and ZP _z are circular holes 32, respectively.
Are the temperature difference and pressure difference at both ends in the longitudinal direction, and the physical properties a, η and ZT _z , ZP _z , T _RA , and F of the fluid are external specifications given externally as capacity conditions to be satisfied by the heat exchanger. It is important that L and Lw _A are determined only by the external specifications as seen from the equations (19) and (21).

On the other hand, R ₀ and Z cannot be determined only by providing external specifications, but have a degree of freedom in which one of the two can be arbitrarily determined. In the following, some important quantities relating to the characteristics of the heat exchanger are shown as a function of these external specifications and R ₀ , but first a mathematical expression of the quantities which are determined only by giving the external specification values.

First, the average flow velocity w _A in the hole of the equation (3) is obtained by the equation (1)
By combining with 9), w _A = (6aZP _z T _RA ) ^1/2 (11ηZT _z ) ^-1/2 (22), which indicates that the amount is determined only by the external specification. As shown in FIGS. 1 and 3, generally, a plurality (N
₁ ) When the total sum of the internal cross-sectional areas of a certain pressure-resistant circular pipe 10 is A, A is also the total sum of the front surface areas of the thin plates 30 installed in the pressure-resistant circular pipe 10. , The sum determines the frontal area of the heat exchanger, and the expression of A is also expressed as A = F / (hw _A ) = (F / h) (6aZP _z T _RA ) ^-1/2 (11ηZT _z ) ^1/2 (23), which indicates that the determination is made only by the external specification. In this equation (23), h is the aperture ratio of the thin plate 30, that is, the ratio of the total cross-sectional area of the large number of circular holes 32 on the thin plate 30 to the front surface area of the thin plate 30 (= the internal cross-sectional area A of the pressure-resistant tube 10). is there. When the area A is composed of one circle, its radius R
₁ is a _{R 1 = (F / (πhw} A)) 1/2 = F 1/2 (πh) -1/2 (6aZP z T RA) -1/4 (11ηZT z) 1/4 (24) However, as shown in FIGS.
Also it is divided into _one. The heat load of the heat exchanger, heat Q exchanged in unit time that is is _{Q = Fρc p ZT z (25} ), the ratio _{E / Q = ZP z / (} ρc of E of the Q of the formula (18) _p ZT _z) (26) represents the percentage of the viscous heating loss for Zenden heat of the heat exchanger. The specific ^{_{S η * / S = (ZP}} z V A / T A) / (c p ZT z / T A) of the ratio E / Q is entropy production losses due to viscous heating S _eta ^* and the entropy rate S = E / Q (27). Where _VA is the volume per mole of fluid. This S _η ^* / S is the second value on the right side of Expression (9).
The term indicates the relative magnitude of the viscous heat loss of the term.
This value can be less than about whether by specifying a smaller pressure loss ZP _z, in the embodiments described later are selected around 1%, the right-hand side of the value of the order of 1% (9) It can be said that the second term (the viscous heating term) is so small that it can be ignored.

Next, an entropy generation loss due to heat conduction in the fluid flowing through the circular hole 32 is calculated. Assuming only the heat conduction in the direction of the radius r, the local entropy generation rate s ^* per unit volume can be expressed as s ^* = κ (∇T) ² / T ^{2 where} T _A is the average temperature in the circular hole 32. = Κ (∂T / ∂r) ² / T _A ² (28), and a value S obtained by integrating this over the entire cross section of the circular hole 32
^* It is calculated ^{S * = ∫2πrs * dr = (} 48/11) πκT RA 2 / T A 2 and (29). Meanwhile total S entropy flowing out circular hole wall _{S = 2πR 0 κ | ∇T /} T | = (2πR 0 κ / T A) | ∂T / ∂a | (r = R 0) = (48/11 ) ΠκT _RA / T _A (30), and the ratio S ^* / S = T _RA / T _A (31) gives the ratio of loss due to entropy generation to total entropy flow. Expression (1) included in this expression (31)
Since the average temperature difference T _RA specified in 5) continues to exist at the outlet of the circular hole 32, the temperature range in which the heat exchanger can heat or cool the fluid should be reduced by this T _RA. Therefore, this T _RA is an important value in the design of the heat exchanger, but it is important that the T _RA be determined if S ^* / S is specified.

[0031] The above is a various amount determined the value just by applying an external specification, since it shows the various amounts affected by the R ₀ as a function of the external specifications and R _0. As described above, R ₀ is an amount whose value can be freely determined independently of the external specification.

The total length Z of the circular holes 32 and the ratio of the length and diameter of the circular holes 32 to Z / (2R ₀ ), that is, the narrow length of the circular holes 32, are as follows: Z = R ₀ ² / L = R ₀ ² (384 ^{_{aηT RA) -1/2 (11ZT z ZP}} z) 1/2 (32) Z / (2R 0) = R 0 (1536aηT RA) -1/2 (11ZT z ZP z) 1/2 (33) circular hole The residence time τ of the fluid in 32 is τ = Z / w _A = R ₀ ² / (Lw _A ) = R ₀ ² (11ZT _z ) (48aT _RA ) ⁻¹ (34), and V = Fτ / h = R ^{_{0 2 (F / h) (}} 11ZT z) (48aT RA) -1 (35) is a volume of the fluid to stay in the heat exchanger. When this V is calculated for both of the two fluids that exchange heat and the sum of the two is obtained, the sum becomes the total volume of the heat exchanger. V is inversely proportional to the fluid property a and the pipe temperature difference T _RA , and the required temperature rise ZT
_Since it is directly proportional to _Z , but also proportional to the square of the hole radius R ₀ , by reducing the hole diameter R ₀ , the heat exchanger volume V can be reduced as desired while keeping the external specifications constant. This theory shows. This fact has not been emphasized in spite of the fact that the Hagen-Poiseuille flow itself has been known for many years, and is the main point of view of the present invention. The total number N of the circular holes 32 provided on the total sum A of the front area of the heat exchanger is N = Ah / (πR ₀ ² ) = R ₀ ^-2 (F / π) (11ηZT _z ) ^1/2 ( 6aZP _z T _RA ) ^-1/2 (36), which is proportional to R ₀ ^-2 . However, N is a total sum of the circular holes 32 through which the fluid 1 flows, and is present in a plurality of pressure-resistant circular tubes 10 in a dispersed manner.

The above analysis is valid only for the Hagen-Poiseuille flow, that is, the laminar flow. The laminar flow exists when the Reynolds number Re defined below is 2000 or less, and Re is It should be noted that above 2000, the above analysis is no longer applicable due to the turbulence of the flow. The expression of Re is: Re = w _A (2R ₀ ) ρ / η = R ₀ (24 aρ ² ZP _z T _RA ) ^1/2 (11η ³ ZT _z ) ^-1/2 = R ₀ (24κρZP _z T _RA ) ^{1 / 2} (11 c _p η ³ ZT _z ) ^-1/2 (37), which is directly proportional to R ₀ after the external specification is determined.

In addition, the total heat transfer area of the heat exchanger is A _H = 2πR ₀ ZN = R ₀ (11FZT _z ) / (24aT _RA ) (38) The heat transfer coefficient of the inner wall of the hole is α = Q / (A _H T _RA ) = R ₀ ⁻¹ (24/11) κ (39), and the Nusselt number Nu in which α is dimensionlessly expressed is Nu = α (2R ₀ /κ)=48/11=4.36 (40 ) and a constant, circular hole resistance coefficient of Hagen Poazoiyu flow 32 is ^{_{λ = 4P z R 0 ρ -1}} W a -2 = (22/3) ηT z R 0 / (ρaT RA) = 256 η ² (ρP _z ) ^-1 R ₀ ^-3 = 64Re ^-1 (41)

Next, a circular hole 32 and a circular hole 3 formed in the thin plate 30 will be described.
The thermal conductivity is a metal portion between the two, but circular holes 3 having a radius R ₀
2 is a thin plate 30 having a radius R ₁ penetrating in the axial direction.
Heat conduction can be approximated by axisymmetric heat conduction circle plate having a uniform heat source b on the entire surface when _{the first} is fitted tightly in a pressure circular pipe 10. The equation governing this heat conduction is κ
The thermal conductivity of the fluid, kappa ₁ the thermal conductivity of the metal part, kappa ₁ ^*
Where ² T = (∂ / ∂r) ² T + (1 / r) (∂T / ∂r) = − b _{(42) b = -ρc p hw} a ZT z / (Zκ 1 *) = (48/11) (κ / κ 1 *) hT RA R 0 -2 (43) solution is axial symmetry of equation (42) considering, comprising the temperature of the inner tube wall temperature distribution _{T = T 1 + (b /} 4) (R 1 2 -r 2) (44) and the quadratic curve when the T _1.

Further, the loss rate S ₁ ^* / S ₁ of entropy generation due to the heat conduction is calculated as T _A ,
It is obtained as follows by an equation in which κ and R ₀ are replaced by T _A1 , κ ₁ and R ₁ respectively.

[0037] ^{_{S 1 * / S 1 = bR}} 1 2 / (8T A1) = hR 1 2 (T A1 κ *) -1 (3κρc p ZT z ZP z T RA) 1/2 (352 η) -1 / ² = T _RA1 / T _A1 (45) In this equation (45), T _RA1 represents the temperature difference T−T ₁ by the area 2πd.
It is weighted by r and averaged, and has a proportional relationship completely corresponding to equation (31). The loss rate S ₁ ^* / S ₁ is R ₁ ²
Therefore, when the value of S ₁ ^* / S ₁ is set to a constant level and an external specification value is given, the value of the radius R ₁ of the inner pipe becomes R ₁ = ((S ₁ ^* / S ₁ ) (8T _A1 / b) ^1/2 = R ₀ ((11/6) (S ₁ ^* / S ₁ ) (T _A1 / T _RA ) (κ ₁ ^* / κ) / h) ^1/2 = R ₀ ((11 / 6) (T _RA1 / T _RA ) (κ ₁ ^* / κ) / h) ^1/2 (46) As a result, the flow rate of the fluid passing through one pressure-resistant circular pipe 10 is also determined, and the number N ₁ of the inner tube that is necessary to ensure that the flow rate F is also _{N 1 = F / (W a} πR 1 2) is determined as (47).

The relationship between the thermal conductivity κ ₁ of the metal part and the effective thermal conductivity κ ₁ ^* in consideration of the heat transfer alienation effect due to the opening is shown in FIG. Focusing on the fact that the two-dimensional heat conduction strictly obeys the Cauchy-Riemann equation, the heat conduction when an infinite number of circular holes are drilled in a square lattice in a flat thin plate that extends infinitely is governed by a kind of elliptic function. can indicate Rukoto, thereby the _{β = R 0 / m (48} ) and placing the kappa ₁ ^* and kappa ₁ ratio when the central interval is 2m adjacent circular hole κ ₁ ^{* /} κ ₁ = (1−β ² ) / (1 + β ² ) (49) This β is the aperture ratio h of the thin plate 30
Since β = (4h / π) ^0.5 (50), κ ₁ ^* / κ ₁ = (1-4 h / π) / (1 + 4 h / π) (51) is obtained. Equation (51) is a high-precision approximation of an exact solution to this heat conduction problem.

The above description has been made with respect to the hole type heat exchanger. However, for the narrow groove type heat exchanger which faces at a minute interval of 2 × ₀ , the following relations are made corresponding to the case of the hole type. Holds. First, the flow velocity distributions corresponding to Expressions (1) to (4) are (u, v, w) = (0, 0, w (x)) (1 ^* ) w (x) = (3/2) w ^{_{A * (1- (x / X}} 0) 2) (2 *) w A * = -P z X 0 2 / (3η) = - ZP z L * / (3η) (3 *) L * = X 0 ² / Z (4 ^* ), and from equation (10) for a circular hole to equation (3)
Temperature distribution T which corresponds to up to 7) (x, z) = T z (z) + T x (x) (12 *) T x = (175/136) T XA (1- (6/5) (x ^{_{/ X 0) 2 + (1/5}} ) (x / X 0) 4) (14 *) T XA = (17/105) P z T z X 0 4 / (aη) (15 *) and is expressed, various representation corresponding to up to equation (37) from equation (19) in the case of a circular tube is ^{_{L * = X 0 2 / Z}} = ((105/17) aηT XA / (ZP z ZT z)) 1/2 ( ^{_{19 *) X 0 = (ZL}} *) = Z 1/2 ((105/17) aηT XA / (ZP z ZT z)) 1/4 (20 *) L * w A * = (35a / 17) T ^{_{XA / (ZT z) (21}} *) W A * = ((35/51) (a / η) (ZP z T XA) / (ZT z)) 1/2 (22 *) Z = X 0 2 ( (105/17) aηT _XA / (ZP _z ZT _z )) ^-1/2 (32 ^* ) τ ^* = Z / w _A ^* = X ₀ ² / (L ^{* _{w A *) = X 0 2}} (17/35) (1 / a) (ZT z / T XA) (34 *) V = Fτ * / h = FX 0 2 (17/35) ZT z / (ahT XA ) (35 ^* ) Re = w _A ^* (2X ₀ ) ρ / η = X ₀ (140 aρ ² zP _z T _XA ) ^1/2 (51η ³ zT _z ) ^-1/2 = X ₀ (140 κρzP _z T _XA ) ^1/2 (51η ³ c _{p z} T _z ) ^-1/2 (37 ^* ), and the equations (6), (9) and (18) hold true in the case of a parallel plate type heat exchanger.

Further formula (28) applies entropy production losses - (31) to the corresponding ^{s * = κ (∇T) 2} / T 2 = κ (∂T / ∂x) 2 / T A 2 (28 * ^{) S * = 2∫s * dx =} (70/17) T XA 2 / (T A 2 X 0) (29 *) S = 2κ (∇T / T) = (2κ / T A) (∂T / ∂x) (at _{x = X 0) = (70/17} ) κT XA / (T a X 0) (30 *) S * / S = T XA / T a (31 *) is satisfied.

When the coefficient value for the above-mentioned narrow groove type is compared with that for the circular hole type, the result is not necessarily advantageous over the circular hole type. Although the narrow groove type has a drawback such as requiring a mechanism for maintaining a constant groove width, it also has an advantage of a simple structure, and is one of the options under the present invention.

In the above-described example, the case where the arrangement positions of the groups of the fine holes (circular holes) 32 of the thin plates 30 are described, but as shown in FIG. 6, the circular holes 32 of the adjacent thin plates 30 are arranged. May be shifted in the direction perpendicular to the + -z direction.

That is, if the positions of the holes and grooves match when viewed from the + -z direction, the fluid flowing through the center of the holes and grooves tends to pass through without performing sufficient heat exchange.
This tendency can be reduced by shifting the positions of the holes and grooves as viewed from the + -z direction. The thermal diffusion distance D is defined as D = (aτ) ^1/2 with respect to the thermal diffusivity a defined by the equation (16) and the residence time τ defined by the equation (34), where D is the diameter of the pore. When the position of the pores is shifted by taking the subsequent n sheets as one cycle when the diameter is smaller than the above, D per one sheet is reduced by n ^-1/2 times,
Since the number of sheets becomes n, D for n sheets becomes n ^{1-1 / 2} = n ^1/2 times, and the same effect as when the diameter of the pores is reduced to 1 / n can be obtained.

FIG. 7 shows the configuration of a nuclear power plant 400 according to one embodiment of the present invention. The nuclear power plant 400 includes the heat exchanger 200 having the above-described configuration, and the helium gas used as the working fluid has both functions of cooling the reactor core and driving the gas turbine.

In FIG. 7, reference numeral 210 denotes a core, 21
5 is a gas turbine, 220 is a generator, 225 is a regenerative heat exchanger, 230 is a pre-cooler, 232 is cooling water, 235 is a first stage compressor, 240 is an intermediate cooler, and 245 is a second stage compressor. It is. Further, in the same figure, a portion surrounded by a dotted line indicates a place where the working fluid exchanges thermal energy or mechanical energy, that is, a thermomechanical element, and a broken line with an arrow indicates a thermal coupling (exchange of heat). , And a dashed line with an arrow indicates mechanical coupling (transfer of mechanical energy).

The thermal efficiency of a heat engine or pump, which originally operates using heat sources of two different temperatures, reaches a theoretical value, ie the highest theoretically possible value, when no entropy is generated through the working fluid. One of the best ways to bring the thermal efficiency of a practical motor closer to this theoretical value is
This is to bring the thermodynamic cycle of the prime mover closer to a cycle called the Ericsson cycle.

The Ericsson cycle consists of two isothermal processes and two
The two equal pressure processes are heat exchange processes (see Mechanical Engineering Handbook). In addition, this isothermal process can be simulated by reducing the temperature change during compression by multi-stage compression with intermediate cooling, that is, by stacking a combination of adiabatic compression and cooling in multiple stages. Isothermal expansion can also be performed.

Such an approximate Ericsson cycle can be used in many applications such as an ice machine, an industrial steam generator, a thermal power plant, and a nuclear power plant. The nuclear power plant 400 shown in FIG. This is an example of a cycle, in which the above-mentioned intermediate cooling 2
Stage compression and regeneration heat exchange are performed. Each of the approximating means of the Ericsson cycle requires a heat exchanger, and the heat exchanger 200 having a large capacity per volume according to the present invention can be used for these.

In the gas turbine nuclear power generation system, the equipment cost of the steam turbine system based on the Rankine cycle used in most of the present thermal power generation and nuclear power generation is extremely high, and it is possible to avoid low thermal efficiency and to reduce the equipment cost. It is promising as one that enables power generation with high thermal efficiency in
For an overview of this method, see the proposal of General Atomics, Inc. of the United States (Nuclear Industry, 40, 12, p31) and the review of this method by Toshiyuki Tanaka of the Japan Atomic Energy Research Institute (Nuclear Industry, 41, 1, p43). Can be.

FIG. 8 shows a thermodynamic cycle of the gas turbine nuclear power generation system in a (P, v) diagram, that is, a (pressure-specific volume) diagram. In this cycle, the adiabatic compression stroke of a normal regenerative gas turbine cycle is replaced by adiabatic two-stage compression with intermediate cooling. The details are as follows.

First, the helium gas passes between the ABs and passes through the core 210 of the nuclear reactor, and receives heat supply from nuclear fuel. BC is an expansion stroke by the gas turbine 215, and the heat between the CDs is the heat radiation in the regenerative heat exchanger 225, and this heat is used for heating between the HAs. Between the DEs, the precooler 230 radiates the helium gas to water, which is a low-temperature heat source. EF and GH are the first-stage and second-stage compression, respectively, and the intermediate FG is intermediate cooling with water in the intercooler 240, and the cycle ends with heating of the HA by the regenerative heat exchanger 225.

[0052] cycle thermal efficiency _{= 1- (T D -T E +} T F -T as the absolute temperature of the thermal efficiency cycle approximate isothermal and isobaric process cycles on the T _i Assuming comprising only the points of the cycle _G ) /
_{(T B -T A) = 59} % and is calculated, General Atomic Company is a 48% thermal efficiency in consideration of various losses are achievable. Incidentally, the maximum temperature T _B of this cycle = 1123 K and the minimum temperature T
_E = thermal efficiency of the Carnot cycle between 303K is Carnot cycle thermal efficiency _{= 1-T E / T B} = 73%.

Now, the heat exchanger according to the present invention can be used in three places of the regenerative heat exchanger 225 between the CDs, the pre-cooler 230 between the DEs, and the intercooler 240 between the FGs in FIG. In the plant proposed by General Atomics, a plate fin type heat exchanger is used, but the use of the heat exchanger according to the present invention reduces the volume of the heat exchanger by about 1/6 as described later. And the same effect can be obtained for the intercooler and the precooler.

The essential point of the regenerative heat exchanger 225 is the formula (4)-
It is calculated as follows using the temperature and pressure conditions of (47) and various points C, D, H, and A shown in FIG. However, the pressure loss ZP _z and borehole temperature difference T _RA are those given as external specification. Numerical values for the high-pressure side corresponding to points H and A are written outside [], and those for the low-pressure side corresponding to points C and D are written inside [].

T _A = 584 K [584] P _A = 7.16 MPa [2.60] ZT _z = 398 deg [398] ρ = 5.90 Kgm ^-3 [2.14] C _p = 5196 Jkg ^-1 K ^-1 [5196] κ = 2.5 × 10 ⁻¹ m ² s ⁻¹ [2.5 × 10 ⁻¹ ] η = 3.09 × 10 ^-5 Pas [3.09 × 10 ^-5 ] G = 668 Kgs ⁻¹ [668] F = G / ρ = 113 m ³ s ^-1 [312] T _RA = 5.0 deg [5.0] ZP _z = 1.70 × 10 ⁵ Pa [1.70 × 10 ⁵ ] S ^* / S = T _RA / T _A = 0.00856 S _η ^* / S = 1.39 × 10 ^{− 2} = 1.39% [3.84] a = 8.15 x 10 ^-6 m ² s ^-1 [2.25 x 10 ^-5 ] aη = 2.52 x 10 ^-10 N [6.95 x 10 ^-10 ] L = 2.55 x 10 ^-8 m [ _{^{4.23 × 10 -8] Lw A =}} 4.47 × 10 -7 m 2 s -1 [1.23 × 10 -6] w A = 17.9ms -1 [29.1] R 1 = 1.03 × 10 -2 m [1.35 × 10 - ² ] N ₁ = 1.88 × 10 ⁴ [1.88 × 10 ⁴ ] is obtained. Here, if the hole radius is set to R ₀ = 2.3 × 10 ⁻⁴ m [3.0 × 10 ⁻⁴ ], Z = 2.13 m [2.13] Z / (2R ₀ ) = 4630 [3550] τ = 0.119 s [0.0732] V = 13.4 m ³ [22.8] N = 0.380 × 10 ⁸ [0.380 × 10 ⁸ ] Re = 1570 <2000 [1210 <2000] As indicated by the above value of V, when the heat exchanger according to the present invention is used in this nuclear power plant, the volume of the regenerative heat exchanger is 13.4 + 22.8 = 36.2 m as a sum of the high pressure side and the low pressure side.
Is a ^3, the volume of the regenerative heat exchanger proposed by General Atomikusu Inc. estimated to be about 200 m ^3, it would be requires only one volume of about 6 minutes by the use of the present invention, largely on the size of the plant To contribute.

Next, the pre-cooler 2 corresponding to between DE in FIG.
The essentials of the heat exchanger 30, that is, the heat exchanger for releasing heat to water, which is a low-temperature heat source of helium gas, are calculated as follows. However ZP _z and T _RA are those given as external specification. The numerical value on the water side is shown in [].

T _A = 346 K [346] P _A = 2.5 × 10 ⁶ Pa [1 × 10 ⁵ ] ZT _z = 79 K [79] ρ = 3.48 Kgm ^-3 [976] C _p = 5196 Jkg ^-1 K ^-1 [4191] κ = 1.7 × 10 ^-1 Wm ^-1 k ^-1 [0.673] η = 2.21 × 10 ^-5 Pas [4.04 × 10 ^-4 ] F = 192 m ³ s ^-1 [0.848] T _RA = 5K [ 5] ZP _z = 1 × 10 ⁵ Pa [1 × 10 ⁵ ] S ^* / S = T _RA / T _A = 0.0145 [0.0145] S _η ^* / S = 0.0125 [0.0629] a = 9.40 × 10 ⁻⁶ m ² s ^-1 [1.65 x 10 ^-7 ] L = 6.78 x 10 ^-8 m [3.84 x 10 ^-8 ] Lw _A = 2.60 x 10 ^-6 [4.56 x 10 ^-8 ] w _A = 38.3 ms ^-1 [1.19] R ₁ = 1.26 × 10 ⁻² m [4.72 × 10 ⁻³ ] N ₁ = 1.00 × 10 ⁴ [1.02 × 10 ⁴ ] R ₀ = 3.0 × 10 ⁻⁴ m [5.38 × 10 ⁻⁴ ] Z = 1.33 m [ 1.33] Z / (2R _O ) = 2220 [1240] τ = 3.47 × 10 ⁻² s [1.12] V = 6.67 m ³ [1.66 × 10 ⁻² ] N = 1.77 × 10 ⁷ [4.29 × 10 ⁶ ] Re = 3620> 2000 [1320 <2000] .

Next, the specifications when the present invention is applied to the intercooler 240 for two-stage compression of helium gas, that is, the heat exchanger corresponding to FG in the cycle diagram shown in FIG. Become.

[0059] _{T A = 345 K [345]} P A = 4.25 × 10 5 Pa [1 × 10 5] ZT z = 178 K [178] ρ = 5.93Kgm -3 [976] C p = 5196JKg -1 K - ¹ [4191] κ = 0.17 Wm ⁻¹ K ⁻¹ [0.662] η = 2.21 × 10 ⁻⁵ Pas [3.94 × 10 ⁻² ] F = 113 m ³ s ⁻¹ [0.851] T _RA = 3K [3] ZP _z = 0.5 × 10 ⁵ Pa [0.1 × 10 ⁵ ] S ^* / S = T _RA / T _A = 0.0087 [0.0087] S _η ^* / S = 0.0239 [0.00477] a = 5.52 × 10 ⁻⁶ m ² s ⁻¹ [1.62 × 10 ⁻⁷ ] L = 6.13 × 10 ⁻⁸ m [4.43 × 10 ⁻⁷ ] Lw _A = 1.06 × 10 ⁻⁶ [3.12 × 10 ⁻⁸ ] w _A = 17.3 ms ⁻¹ [0.0704] R ₁ = 6.70 × 10 ^-3 m [9.06 × 10 ^-3 ] N ₁ = 4.63 × 10 ⁴ [6.23 × 10 ⁶ ] R ₀ = 2.0 × 10 ^-4 m [5.38 × 10 ^-4 ] Z = 0.653 m [0.653] Z _{/ (2R O) = 1630 [} 607] τ = 4.57 × 10 -2 s [9.28] V = 6.09m 3 [11.3] N = 5.20 × 10 7 [1.33 × 10 7] Re = 1860 <2000 [461 <2000 ] In the present invention In accordance with the above-mentioned specification values, the heat exchanger 200 according to the above-described gas turbine nuclear power generation system has a regenerative heat exchanger 225, a pre-heat exchanger 230, and an intercooler 240.
By using the heat exchanger at a location, the volume of the heat exchanger can be reduced to a fraction of that of a conventional plant.

[0060]

As described above, according to the present invention,
It is possible to provide a heat exchanger and a heat machine having extremely large heat exchange capacity per unit volume, which has been conventionally difficult to achieve.

[Brief description of the drawings]

FIG. 1 is a diagram showing a main configuration of a heat exchanger according to an embodiment of the present invention.

FIG. 2 is a diagram showing a cross-sectional configuration of the heat exchanger of FIG.

FIG. 3 is a diagram illustrating a main configuration of a heat exchanger according to another embodiment of the present invention.

FIG. 4 is a diagram showing a cross-sectional configuration of the heat exchanger of FIG. 3;

FIG. 5 is a diagram showing a main configuration of a heat exchanger according to another embodiment of the present invention.

FIG. 6 is a diagram showing a main configuration of a heat exchanger according to another embodiment of the present invention.

FIG. 7 shows a configuration of a nuclear power plant according to one embodiment of the present invention.

FIG. 8 is a diagram showing a thermodynamic cycle of the nuclear power plant of FIG. 7;

FIG. 9 is a diagram showing a configuration of a main part of a conventional multi-tube heat exchanger.

FIG. 10 is a diagram showing a cross-sectional configuration of the multi-tube heat exchanger of FIG. 9;

FIG. 11 is a diagram showing a configuration of a main part of another conventional multi-tube heat exchanger.

FIG. 12 is a diagram showing a cross-sectional configuration of the multi-tube heat exchanger of FIG. 10;

[Description of Signs] 10: Pressure-resistant tube 11: Space inside pressure-resistant tube 13: Space outside pressure-resistant tube and inside pressure-resistant tube 15: Thin annular pressure-resistant tube 20: Pressure-resistant Tube 30 Thin plate 31 Annular groove 32 Fine pore 34 Gap separating thin plate 100 Heat exchanger F1 Fluid 1 F2 Fluid 2 P1 ... pressure of fluid 1 P2 ... pressure of fluid 2 200 ... heat exchanger 210 used for helium gas turbine nuclear power generation system ... core 215 ... gas turbine 220 ... generator 225 ... regeneration Heat exchanger 230 230 Precooler 232 Cooling water 235 First stage compressor 240 Intercooler 245 Second stage compressor 400 Helium gas Turbine nuclear power system

 ──────────────────────────────────────────────────続 き Continuation of the front page (71) Applicant 592073101 IBM Japan, Ltd. 3-2-1-12 Roppongi, Minato-ku, Tokyo (71) Applicant 000004237 NEC Corporation 5-7-1 Shiba, Minato-ku, Tokyo No. 1 (72) Inventor Eiichi Goto 3-9-305, Tsujido East Coast, Fujisawa-shi, Kanagawa Prefecture (72) Inventor Susumu Kase 710-67, Hirohakamacho, Machida-shi, Tokyo 2-1 Hirosawa, Ichino RIKEN

Claims (5)

[Claims] 1. Two fluids that exchange heat with each other are separated by a pressure-resistant wall made of a material that can withstand a pressure difference between the two fluids, and the two fluids flow in opposite directions to each other. Among the fluids, a thin plate of a material having a high thermal conductivity that is substantially in contact with the pressure-resistant wall and extends in a direction perpendicular to the direction in which the fluid flows is disposed in at least one of the fluids, and the fluid passes through the thin plate In this way, a group of pores or grooves is formed to reduce the thermal resistance between the fluid and the thin plate, and a plurality of the thin plates are arranged side by side with a small gap in the flow direction of the fluid to form a flow direction of the fluid. A heat exchanger that reduces heat conduction to the pores and the fluid pressure in the groups of pores or grooves. 2. The pressure-resistant tube, wherein the pressure-resistant wall is at least one pressure-resistant tube through which one of the two fluids flows inward and the pressure-resistant tube through which the other fluid flows inward. The heat exchanger according to claim 1, wherein the heat exchanger is constituted by a tube wall. 3. The heat exchanger according to claim 2, wherein the pressure-resistant wall is formed by disposing a member having a closed curve in cross section and having a height of the small gap between the thin plates. vessel. 4. The heat exchanger according to claim 1, wherein at least a part of the pores or grooves of the adjacent thin plates are arranged at positions shifted from each other. 5. An entropy generation loss by bringing a working gas into thermal contact with two high and low heat sources, each having a substantially constant temperature, and performing an isothermal compression, an isothermal expansion and an isobaric heat exchange on the working gas. A heat machine for approximately realizing an Ericsson cycle without any heat, comprising the heat exchanger according to any one of claims 1 to 4.