U.S. patent number RE33,528 [Application Number 07/371,663] was granted by the patent office on 1991-01-29 for microtube-strip heat exchanger.
Invention is credited to F. David Doty.
United States Patent |
RE33,528 |
Doty |
January 29, 1991 |
**Please see images for:
( Certificate of Correction ) ** |
Microtube-strip heat exchanger
Abstract
A new approach to the theory of heat exchanger optimization is
presented which shows the advantages of using low Reynolds and
Nusselt numbers and low flow velocities along with a novel design,
the microtube-strip (MTS) counterflow heat exchanger. The MTS
exchanger in the preferred embodiment consists of a number of small
modules connected in parallel. Each module typically contains eight
rows of one hundred tubes, each of 0.8 mm outside diameter and 0.16
m length. The tubes are metallurgically bonded via the diffusion
welding technique to rectangular header tube strips at each end.
Caps suitable for manifolding are welded over the ends. Cages are
provided to cause the shell-side fluid to flow in counterflow
fashion over substantially all of the tube length, and suitable
manifolds are provided to connect the modules in parallel. This
design results in the highest power densities of any known design
for single phase exchangers. Although the MTS exchanger of the
present invention is specifically optimized for applications not
involving phase changes in the working fluid, the essential
concepts and features of this invention can also be advantageously
used in applications involving change of phase.
Inventors: |
Doty; F. David (Columbia,
SC) |
Family
ID: |
27005466 |
Appl.
No.: |
07/371,663 |
Filed: |
June 23, 1989 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
Issue Date |
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243048 |
Sep 9, 1988 |
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Reissue of: |
700125 |
Feb 11, 1985 |
04676305 |
Jun 30, 1987 |
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Current U.S.
Class: |
165/158; 165/173;
165/160 |
Current CPC
Class: |
F28F
9/18 (20130101); F28F 9/26 (20130101); F28D
7/1653 (20130101); F28F 9/013 (20130101); F28F
2280/02 (20130101); F28F 2260/02 (20130101) |
Current International
Class: |
F28F
9/013 (20060101); F28F 9/26 (20060101); F28F
9/04 (20060101); F28F 9/007 (20060101); F28D
7/00 (20060101); F28D 7/16 (20060101); F28F
9/18 (20060101); F28F 009/02 () |
Field of
Search: |
;29/157.4
;165/158,159,160,172,173,177,905,910,168,185 ;228/193 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
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965803 |
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Jun 1957 |
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DE |
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2120544 |
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Nov 1972 |
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DE |
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2422168 |
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Nov 1975 |
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DE |
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1130461 |
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Oct 1956 |
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FR |
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2264620 |
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Oct 1975 |
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FR |
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371608 |
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Apr 1932 |
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GB |
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1141102 |
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Jan 1969 |
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GB |
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1491914 |
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Nov 1977 |
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GB |
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Other References
"Heat Exchangers: Design and Theory Sourcebook", by Afgan et al.,
(1974), Chapter 35, (no month provided)..
|
Primary Examiner: Kamen; Noah P.
Attorney, Agent or Firm: Brumbaugh, Graves, Donohue &
Raymond
Parent Case Text
.Iadd.This application is a continuation of application Ser. No.
243,048, filed on 9/9/88, now abandoned, which is a reissue of
application Ser. No. 700,125, filed on 2/11/85, U.S. Pat. No.
4,676,305. .Iaddend.
Claims
What is claimed is:
1. A gas-gas laminar-flow heat exchanger module which
comprises:
a plurality of heat transfer augmentation-free corrosion resistant,
precision, hardened, metallic tubes .[.arrayed in at least four
parallel disposed planar rows of at least forty tubes per row.].
.Iadd.arrayed in a plurality of rows of tubes.Iaddend.;
a first rectangular header strip interference press fit and
diffusion welded to one end of each of said tubes;
a second rectangular header strip interference press fit and
diffusion welded to the other end of each of said tubes;
first manifold means metallurgically connected to said first
rectangular header strip for defining a gas inlet flow path into
said one end of each of said tubes;
second manifold means metallurgically connected to said second
rectangular header strip for defining a gas outlet flow path from
said other end of each of said tubes;
means disposed externally of said tubes for defining a counterflow
flow-path of heat exchanger gas over substantially the entire
length of the external surfaces of each of said tubes from within
the vicinity of said other end of each of said tubes to within the
vicinity of said one end of each of said tubes;
each of said tubes having an outside diameter of less than 3
mm;
each of said tubes having a length which is sufficient to allow for
fully developed laminar flow and which is less than 300 times and
outside diameter of each of said tubes; and
said plurality of tubes within each of said rows being laterally
spaced by a center-to-center distance of from 1.3 to 2.8 times the
outside diameter of each of said tubes.
2. A module according to claim 1 wherein said tubes have been
produced from high tensile strength alloy metal.
3. A module according to claim 2 wherein said high strength metal
alloy is stainless steel.
4. A module according to claim 1 wherein said rectangular header
strips are reinforced against high gas pressure by reinforced
plates.
5. A module according to claim 1 wherein said tubes are supported
at one or more locations by stiffening wires welded between said
rows.
6. A module according to claim 1 wherein said outside diameter of
each of said tubes is approximately 0.8 mm.
7. A module according to claim 6 wherein said length of each of
said tubes is approximately 0.16 m.
8. A module according to claim 1 wherein said array of tubes
comprises .[.eight.]. .Iadd.at least 4 parallel disposed planar
.Iaddend.rows of said tubes .[.with from 40 to 200.]. .Iadd.with at
least 40 .Iaddend.tubes per row.
9. A module according to claim 1 wherein said tubes are disposed
within said rectangular header strips through means of a 0.3% to 5%
interference press fit.
10. A module according to claim .[.1.]. .Iadd.8 .Iaddend.wherein
the distance between said parallel rows is equal to 0.866 times
said center-to-center tube distance.
11. A module according to claim 1 wherein said means external of
said tubes for defining said counterflow flow-path of said heat
exchanger gas comprises a cage annularly surrounding said array of
tubes.
12. A module according to claim 1 further comprising a plurality of
said modules defined by said tubes, said first and second header
strips, and said first and second manifold means, vertically
stacked together; and third and fourth manifold means, respectively
connecting together said sets of first second manifold means of
each of said modules, for supplying said tube-side gas inlet and
outlet paths from said modules; whereby said modules and said third
and fourth manifold means define a heat exchanger module block.
13. A module according to claim 12 further comprising pressure
vessel means for housing said block and thereby defining a heat
exchanger module tank.
14. A module according to claim 13 further comprising cage means
surrounding and enclosing said first and second manifold sets and
said third and fourth manifold means for defining a counter-flow
flow-path of heat exchanger gas externally of said plurality of
tubes.
15. A module according to claim 14 wherein said cage means are
disposed internally within said pressure vessel means.
16. A module according to claim 15 further comprising expansion
joint means defined between at least one of said cage means and
said pressure vessel means for relieving axial thermal stresses.
.Iadd.17. A module according to claim 8 wherein said array of tubes
comprises eight rows of said tubes with from 40 to 200 tubes per
row. .Iaddend.
Description
BACKGROUND OF THE INVENTION
The field of this invention is heat exchangers, and more
particularly, counterflow, modular, shell-and-tube-type exchangers
for single phase fluids with no heat transfer augmentation
means.
PRIOR ART
The result of four decades of industrial and commercial interest in
heat exchangers has seen a proliferation of specialized devices and
manufacturing techniques that offer some advantages in special
applications. The present invention is based on a radical departure
from conventional heat exchanger design guidelines in several
distinct areas. As a result, the design differs in a number of
ways, but the most significant innovative feature is the most
subtle and is not apparent without a detailed theoretical
explanation. This most important feature is its size. This change
represents such radical departures from conventional practice in
typical Nusselt and Reynolds numbers as to make reference to prior
art of limited value. Nonetheless, for reference value and
completeness, a brief synopsis of the prior art is presented.
Numerous examples of modular, counter-flow shell and tube
exchangers can be found in the patent literature, one of the
earlier examples being Rossi's bi-directional flow design, U.S.
Pat. No. 2,839,276, with its advantages of reduced thermal
stresses. A more typical recent design is that of Baumgaertner et
al, U.S. Pat. No. 4,221,262, which offers some construction
advantages over earlier designs due to the reduced complexity of
its basic modules. Quite atypical and impractical, but of revelance
on account of its general system appearance, is Giardina's U.S.
Pat. No. 4,253,516, with its huge box-car sized modules.
Jabsen et al in U.S. Pat. No. 4,289,196 and Culver in U.S. Pat. No.
4,098,329 employ unique heading and manifolding systems in attempts
to achieve higher power densities in modular systems. Cunningham et
al give attention to hot corrosive problems in U.S. Pat. No.
2,907,644. Lustenader recognizes the problem of axial conduction
losses in U.S. Pat. No. 3,444,924, a problem obviously not
understood by most heat exchanger design engineers.
Corbitt et al address the problem of vortex induced resonances in
cross flow exchangers, U.S. Pat. No. 2,655,346, and solve it via
the strategic positioning of baffles. Scheidl uses a tube support
grid to solve these problems in U.S. Pat. No. 3,941,188.
Bays, U.S. Pat. No. 2,537,024, and Malewicz, U.S. Pat. No.
3,452,814, give several examples of heat flow augmentation, which
is easily shown to be of negative value in a gas-gas heat exchanger
optimized according to the present invention.
Various well-known joining techniques include Cottone and
Sapersteine's use of special braze alloys, U.S. Pat. No. 4,274,483,
Olsson and Wilson's cold pressure welding, U.S. Pat. No. 4,237,971,
Hardwick's explosive welding, U.S. Pat. No. 3,717,925, Brif and
Brif's expanded tubes, U.S. Pat. No. 4,239,713, and the related
technique of Yoshitomi et al, U.S. Pat. No. 4,142,581. More closely
related to the diffusion technique of the present invention is the
press-fit method of Nonnenmann et al, U.S. Pat. No. 4,159,741, and
the compression method of Takayasu, U.S. Pat. No. 3,922,768.
However, these techniques as described fall short of producing a
high integrity metallurgical bond. Lord's U.S. Pat. No. 4,528,733
describes a joining technique suitable for applications in which
the header is made of a material which undergoes a phase change
that is accompanied by an abrupt change in dimension. Mattioli et
al, U.S. Pat. No. 3,849,854, describe a method of effecting
diffusion welds via induction heating followed by electromagnetic
compression that is suitable for large, accessible joints. Woods,
U.S. Pat. No. 2,298,996, describes a method of hard brazing
aluminum and copper alloy 6 mm tubes, extending beyond their
rectangular headers and expanding into a polygonal shape so as to
reduce tube side pumping losses in turbulent flow applications,
into rectangular headers, while Troy, U.S. Pat. No. 3,782,457,
describes the use of 2 mm tubes in an annular header with heat
transfer augmentation.
Frei's U.S. Pat. No. 4,295,522, employing glass tubes and silicone
casting resins, shows a striking resemblance from a non-scaled
perspective between his basic tube assembly modules and the present
invention. Furthermore, the tube sizes employed therein also show
progressive design traits, being about 6 mm in diameter rather than
the customary 1.5 cm to 2.5 cm employed in all other above
referenced patents. However, Frei's design, aside from temperature
and pressure limitations imposed by the choice of materials,
suffers from the inefficiencies inherent in a cross-flow design, as
necessitated by his manifolding scheme.
The use of small diameter tubes-O.D. less than about 3 mm-has been
predominantly limited to two-phase cross-flow systems. Early
examples may be found in aircraft oil-coolers such as that by
Anderson, U.S. Pat. No. 2,449,922, and the later art. The only
apparent application involving the use of tubes under 1 mm O.D. is
that of Christen et al, U.S. Pat. No. 4,098,852, which employs
osmotic or ultrafiltering polymeric tubes and vaporizing liquids.
Christen's patent also utilizes the shortest tubes found in the
prior art in counterflow exchangers, such length being only about
0.6 m, compared to the more typical length of about 5 m. Roma's
U.S. Pat. No. 4,030,540 is cited as a typical example of prior art
design guidelines that often resulted in such unsound objectives as
attempting to maximize tube length, whereas the correct objective
is always to minimize tube length while satisfying several
additional criteria.
Some useful related theoretical background materials may be found
in two of my earlier patents, although these inventions are quite
remote from the present invention; U.S. Pat. No. 4,321,962
describes a solar energy heat exchanger and storage system; and
U.S. Pat. No. 4,456,882 describes a high-speed turbine-driven
air-bearing-supported sample spinner.
SUMMARY OF THE INVENTION
The present invention, the microtube-strip (MTS) counterflow heat
exchanger, in the preferred embodiment consists of a number of heat
transfer augmentation-free small modules connected in parallel.
Each module typically contains eight rows of one hundred tubes,
each of 0.8 mm outside diameter and 0.16 m length. The tubes are
metallurgically bonded to rectangular header tube strips at each
end. Caps suitable for manifolding are welded over the ends. Means
are provided to cause the shell-side fluid to flow in counterflow
fashion over substantially all of the tube length, and suitable
manifolds are provided to connect the modules in parallel. Power
capacity per unit volume per unit temperature difference of the MTS
exchanger exceeds that of prior art typical designs by a factor of
ten to 1000. Power capacity per unit cost per unit temperature
difference of the MTS exchanger may exceed that of prior art
designs by a factor as large as 10 in some cases. Flow conditions
in the microtubes are fully laminar and extremely subsonic.
Various other objects, features, and attendant advantages of the
present invention will be more fully appreciated as the same
becomes better understood from the following detailed description,
when considered in connection with the accompanying drawings,
wherein:
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is an isometric drawing of an MTS sub-assembly.
FIG. 2 is a plane section view of an MTS header.
FIG. 3 is an isometric drawing of an MTS module.
FIG. 4 illustrates two reinforcement techniques for MTS modules
operating with high tube-side pressure.
FIG. 5 is an isometric drawing of a plurality of MTS modules
manifolded together in parallel to form an MTS block.
FIG. 6 illustrates an MTS block enclosed in a pressurized
vessel.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
The Heat Exchange Power
The usual approach to heat transfer problems is to begin with the
following equation:
where P.sub.h is the heat transfer power (W), h is the heat
transfer coefficient (W/m.sup.2 K), A is the surface area
(m.sup.2), and T.sub..delta. is the temperature difference (K). The
problem then is to determine suitable expression for h under
various conditions. Unfortunately, most engineers, after looking at
equation (1), thereafter tacitly assume that the heat exchange
power is proportional to the total surface area. It is this
erroneous underlying assumption that has virtually stagnated
progress in signal phase heat exchanger design for four decades.
The often overlooked fact is that the complicated heat transfer
coefficient, h, is always inversely dependent on a characteristic
dimension of the heat exchanger, often in such a way that P.sub.h
increases only as the square root of the area. In some case,
P.sub.h may be independent of certain changes in the area, and in
other cases P.sub.h may actually be decreased by an increase in the
area.
Consider first, for example, the tube-bundle heat exchanger with
high turbulent gas flowing through the tubes, which are bathed in a
constant temperature fluid. The conventional approach is to write
the heat transfer coefficient in terms of the dimensionless Nusselt
number, Nu.
where d is the inside diameter (m) of the tubes and k is the
thermal conductivity (Wm.sup.-1 K.sup.-1) of the gas. The Nusselt
number is then expressed in terms of two additional dimensionless
groups, the Prandtl number, Pr, and the Reynolds number, Re.
where C.sub.p is the constant pressure specific heat (J/kgK), and
.mu. is the dynamic viscosity (kgm.sup.-1 s.sup.-1).
where .rho. is the density of the gas (kg/m.sup.3), v is the mean
velocity of the gas (m/s), and G is the mass flow rate per tube
(kg/s). Then, for highly turbulent flow it can be demonstrated
that,
Combining equations (2) through (5) gives the following expression
for the heat transfer coefficient.
Thus, for a given turbulent mass flow rate through a bundle of
tubes of length L, the heat exchange power of equation (1) is
proportional to the length, and inversely proportional to the 0.8
power of the diameter. Hence, increasing the area by increasing the
tube diameter actually decreases the heat exchange power, and the
advantages of short tubes of small diameter are readily
apparent.
Now consider the case of a tube-type counterflow laminar-flow heat
exchanger with center-to-center tube spacing equal to 1.4 times the
outside diameter of the tubes and twice the inside diameter.
Further assume that the thermal conductivity of the tube material
is much greater than the thermal conductivity of the fluids. For
this case, it can be shown that the heat exchange power is
independent of the tube diameter, and is given by the following
expression: ##EQU1## where n is the number of tubes, k.sub.1 is the
thermal conductivity of the inner fluid, and k.sub.2 is the thermal
conductivity of the outer fluid.
From the above discussion it appears that there is little utility
in evaluating a heat exchanger in terms of a heat exchange
coefficient of dimensions Wm.sup.-2 K.sup.-1 as is customary in the
professional and patent literature. Rather, a more useful
characterization is the total effective flow length, nL. By
defining nL as the quotient of P.sub.h and a generalized function
of k.sub.1 and k.sub.2, one arrives at a useful method of comparing
diverse designs-including those which incorporate heat transfer
augmentation means such as extended or roughened surfaces.
Power Losses
The power, P.sub.p1 required to pump a fluid through the heat
exchanger tubes is given by:
where .DELTA.p is the pressure drop (Pa) through the exchanger,
A.sub.f is the frontal fluid area (m.sup.2), and v is the mean
fluid velocity (m/s).
For simplicity, consider the case of laminar fluid flow through
long, smooth tubes. This condition exists for Reynolds numbers, Re,
below 2000. The pressure drop, .DELTA.p, in a fluid flowing through
a tube under laminar conditions is given by:
thus:
The shell-side pumping power loss, P.sub.p2, required to pump fluid
around the tubes can be expressed by a similar equation:
where the gas parameters .mu. and v now refer to the external gas,
and the coefficient f is a complicated function of tube diameter
and spacing. For the standard hexagonal-close-pack pattern with the
distance between tube centers equal to 1.4 times the tube outside
diameter, f is approximately equal to 200.
In addition to the pumping power loss, there is another internal
loss mechanism present in counterflow exchangers which may limit
the thermodynamic efficiency: the axial thermal conduction power of
the tube metal, P.sub.m.
where w is the wall thickness of the tubes (m), k.sub.m is the
thermal conductivity of the tube metal (Wm.sup.-1 K.sup.-1),
T.sub.H is the mean temperature at the hot end, and T.sub.C is the
mean temperature at the cold end.
Optimization
The power available, P.sub.i, from the input gas is:
where C.sub.p is the constant pressure specific heat (J/kgK), and G
is the mass flow rate (kg/S) and is equal to .rho.A.sub.f v. The
waste heat, P.sub.o is
where T.sub..delta. is, as defined earlier, the mean temperature
difference between the counterflowing gases.
Accounting for the losses, the available heat exchange power,
P.sub.E, is ##EQU2## Equating input and output power gives, under
steady-state conditions, the following:
The above equations can now be solved for T.sub.67 using the
definition of mass flow rate and assuming w=d/3. ##EQU3##
This equation depends only on three geometric variables, n, L, and
d, and is reasonably valid for tube-type counterflow laminator heat
exchangers, subject to several above mentioned assumptions. One can
now calculate the power losses and the available heat exchange
power for a given set of thermodynamic and geometric conditions.
The design can be optimized via the linear programming technique of
maximizing an objective function, F.sub.c, such as the
following:
where a and b may have values of 10 and 2 respectively. It becomes
apparent after exercising a linear programming technique on
equation (18) that by giving proper attention to minimizing costs
associated with tube cutting and end preparation, header hole
punching, and tube assembly and insertion techniques, optimized
high power single phase heat exchangers take on a totally new
appearance. They consist of hundreds or perhaps thousands of small
modules, each of which consists of hundreds of small, short tubes.
Reynolds numbers inside the microtubes for these optimized designs
range from 25 to 400, compared to the more common prior art values
of 10,000 to 100,000; and Nusselt numbers are less than 5, compared
to the typical prior art values of 20 to 400. The result is fully
developed laminar flow, tube side and shell side, and flow
velocities below one tenth the speed of sound.
Alternatively one may choose as objective function F.sub.v such
that
Astoundingly, this function is unbounded. In other words, it is
theoretically possible to increase the power-to-volume ratio
without limit, without increasing pumping losses, if one can reduce
the tube diameter and length and increase the number of tubes
without limit. Of course, the above equations cease to be valid
under molecular flow conditions.
The Tubing
Current practice in tube-type counterflow exchangers generally uses
induction-welded steel, copper, or aluminum tubes of about 3 mm to
25 mm diameter with lengths ranging from 0.5 to 6 m and wall
thickness of about 0.25 mm to 3 mm. However, recent advances in
high speed laser welding and super-hard die technology now make it
possible to produce very small stainless steel hypodermic tubing at
very low production costs-less than $0.10 per meter. It is thus
practical to consider the use of tubing with an outside diameter of
less than 1 mm.
Reducing the tubing diameter by a factor of 10 requires the length
to be reduced by a factor ranging from 30 to 100 while the number
of tubes is increased by a similar factor in order to maintain the
same heat exchange power and pumping power loss. However, the total
volume of the heat exchanger is likewise reduced. Furthermore, the
maximum internal pressure rating of the heat exchanger will
probably be increased due to an increase in the relative wall
thickness.
To facilitate rapid assembly of large numbers of small tubes, it is
necessary to depart from the disc shaped tube header sheet normally
used in heat exchangers and instead use a rectangular tube header
sheet or strip. Furthermore, to minimize tube flexing and to reduce
support requirements, it is also desirable to keep the tube length
relatively short. This will also insure that the buckling strength
of the tubes is large enough to permit pressing them into the tube
strip. Moreover, it will raise the transverse acoustic resonance
modes of the tubes thereby making it more difficult to excite such
resonances by turbulence. Also, equations 10 and 11 show that
reducing the tube length will reduce the pumping power losses.
The maximum practical tube length for high-modulus, high strength
alloys such as strain-hardened stainless steel or
precipitation-hardened superalloys is about 300 times the outside
diameter of the tubes, while the maximum practical length for
copper or aluminum tubes is about half that amount. There are
several additional reasons for preferring stainless steel or
superalloys over the more common heat exchanger metals: (1) They
have very low thermal conductivity which may make them easier to
laser weld, but most importantly reduces the internal axial
conduction loss mechanism, P.sub.m, in the counterflow exchangers;
(2) Their high tensile strength allows higher working pressures;
and (3) Their corrosion and high temperature strength properties
are essential in many applications.
Welding and Manifolding
The key to the current invention is the recognition of the
advantage of using small diameter tubing in very short lengths. Its
implementation depends on technological breakthroughs in the
assembly, welding, and manifolding of these tubes. Since the tubes
are very short, it is necessary to resort to narrow modules in
order that counterflow conditions be established over the major
portion of the tube length and also to reduce the inefficiencies
due to non-uniform flow. While a cross-flow arrangement could be
used to circumvent the above mentioned non-uniform flow problems,
such as arrangement would greatly reduce the thermodynamic
efficiency. The counterflow-serial-crossflow arrangement commonly
used in large installations allows somewhat higher efficiency than
the crossflow arrangement but at increased pumping losses. Hence,
the most satisfactory solution is that of narrow modules of four to
twenty rows of tubes.
The extremely small size of the tubes makes almost all types of
conventional welding methods impractical, and the extremely large
number of tubes eliminates most types of individual tube welding
techniques, probably including automated electron beam and laser
techniques because of process control problems arising from thermal
expansion during the welding operations. Two viable options for the
tube-to-strip welds are fluxless brazing and diffusion welding. A
wide variety of conventional welding techniques are suitable for
the rest of the welds.
In the fluxless brazing technique, the braze metal is plated onto
the inside of the holes and onto the outside of the tubes prior to
assembly. After assembly, the complete module is heated in vacuum
or inert atmosphere to the liquidus temperature of the braze metal.
This method is not suited for very high temperature exchangers.
Diffusion welding can be accomplished if the tube diameter and hole
size can be held to very tight tolerances. The use of hardened
tubes and annealed tube strips then makes it possible to press the
tubes into slightly undersized holes. With proper attention to
surface quality and a minimum of 0.3% interference press fit, a
strong metallurgical bond can be formed simply by heating the
assembly to about 0.8 times the absolute melting temperature (K).
This method is suitable for the highest temperatures and all
alloys.
Corrosion
In many cases, heat exchangers must operate in severely corrosive
environments. Under these conditions, it is no longer theoretically
possible to increase the power-to-volume ratio without limit. The
current state-of-the-art in corrosion resistant alloys, such as
Nimonic 81, limits the minimum wall thickness of about 50 microns
for moderately corrosive environments and about 200 microns for
severely corrosive environments. Although the tubes themselves are
too small to make coatings or laminations practical with current
technology, such measures may be applied to the tube strips and to
the manifolds for economy of materials or to achieve combined high
temperature strength and hot corrosion resistance.
Thermal Response Time
In many applications, particularly in the case of mobile gas
turbines, fast response times are necessary for efficient
operation. Currently, a typical 2000 KW gas turbine may have a
mechanical response time of one minute, but the thermal response
time of the heat exchangers incorporated into the system may be ten
hours. Increasing the power-to-mass ratio of the heat exchanger by
the amount possible with the MTS design could reduce the thermal
time constant to less than one minute. Such a dramatic reduction in
mass and thermal time constant opens up many new applications in
all areas of transportation-especially aerospace.
High Pressure Applications
In many applications, for example, in recuperators used in closed
cycle gas turbines, it is necessary to maintain both the internal
(tube-side) and the external (shell-side) fluids at high pressure.
The narrow width of the tube header strip makes this design well
suited to high tube-side pressures. When high shell-side pressures
are required, the entire heat exchanger must be enclosed in a
pressurized containment vessel. The small size of the heat
exchanger simplifies this task.
DETAILED DESCRIPTION OF THE DRAWINGS
The basic unit in the MTS heat exchanger is the MTS sub-assembly as
illustrated in FIG. 1. It consists of typically eight rows of heat
transfer augmentation free microtubes 1 with typically 40 to 200
microtubes in each row. The microtubes are diffusion welded into
precision MTS header strips 2 at each end. The diffusion welding is
accomplished by using ultra precision, diamond-die-reduced, laser
welded hard drawn tubing for the microtubes, and precisely
machining the holes in the annealed header strip to a size at least
0.3% smaller but not more than 5% smaller than the tubing outside
diameter. A combination of techniques may be required to produce
the precision holes in the header strips, including feinblanking,
electrochemical machining, and reaming. The diffusion welds are
accomplished by (1) insuring that the tubes and holes have
thoroughly cleaned, oxide-free surface prior to assembly, (2)
maintaining a minimum of 0.3% interference press fit, (3) heating
the sub-assembly in an inert atmosphere or vacuum to a temperature
of approximately 80% of the absolute melting temperature of the
tube or header strip alloy, whichever is lower.
FIG. 2 illustrates the recommended HCP (hexagonal close pack) hole
pattern for the MTS header strip 2. The distance between rows is
equal to 0.866 times the distance between tube centers, TC, which
is generally about 1.3 to 2.8 times the O.D. of the sample tubes
1.
FIG. 3 illustrates the basic counterflow MTS module. It includes a
semi-cylindrical cap 3 welded to each header strip. Care is taken
to assure that the header strip 2 is no wider than is necessary to
accommodate the microtubes 1 and the relatively thin walled cap 3
so that the MTS modules may be mounted closely in parallel.
Tube-side manifold ports 4 are provided on each cap 3. A cage 5
closely surrounds the MTS sub-assembly, except near each header
strip, forcing shell-side fluid 6 to enter around the periphery of
the MTS sub-assembly near one end and to exit in like fashion at
the other end. Tube-side fluid 7 enters the tube-side manifold
ports 4 at the end at which the shell-side fluid exits, and it
exits in like manner at the opposite end.
In certain applications, extremely high tube-side pressures,
perhaps combined with very high temperatures, may require
additional support of the flat header strip 2, to prevent bowing of
this surface. This additional support may be provided as shown in
FIG. 4 by diffusion welding a reinforcement plate 8 similar to the
header strip 2 a short distance from it. Alternatively, the
required support may be provided by the microtubes 1 if they are
supported in such a way to prevent their buckling. This may be
accomplished by bonding, preferably by projection welding,
stiffening wires 9 crosswise between the rows of microtubes 1. By
staggering or offsetting the location of adjacent stiffening wires
9, the effect on fluid flow is generally made negligible.
FIG. 5 illustrates the parallel manifolding of several MTS modules
to form an MTS block. Individual fluid ports 4 are connected to a
tube-side manifold 10 at each end. The manifold cages 11 in
cooperation with the MTS module cages 5 form the shell-side sealed
region. Tube-size fluid may exit at tube-side manifold port 12
while shell-side fluid may enter at manifold cage port 13. The MTS
modules are supported by the headers, with adequate clearance space
between the adjacent caps to permit the required shell-side flow 6
between caps with acceptable pressure drop. Typical MTS blocks may
include four to fifteen MTS modules in parallel, and typical high
power installations may include hundreds of such MTS blocks further
manifolding in parallel.
FIG. 6 depicts an MTS block mounted inside a pressure vessel 14
forming an MTS tank for applications requiring high shell-side
pressures. Pressure equalizing vents 15 are required to equalize
mean static pressure components on the flat surface of the MTS
cages 5 and manifold cages 11. The dynamic pressure components
arising from the shell-side fluid pressure drop through the MTS
block must be kept relatively small to prevent excessive deflection
of the flat surfaces. Expansion joints 16 are required at one end
to relieve axial thermal stresses. Suitably sealing flanges 17 and
18 are provided to permit convenient assembly of the containment
vessel 14 and adequate sealing around the ports 12 and 13. Suitable
radial support for the MTS block within the vessel is required at
the end which includes the expansion joints 16.
Although this invention has been described herein with reference to
specific embodiments, it will be recognized that changes and
modifications may be made without departing from the spirit of the
present invention. All such modifications and changes are intended
to be included within the scope of the following claims.
* * * * *