U.S. patent application number 16/316478 was filed with the patent office on 2020-06-11 for corrugated tube-in-tube heat exchangers.
This patent application is currently assigned to Stone Mountain Technologies, Inc.. The applicant listed for this patent is Stone Mountain Technologies, Inc.. Invention is credited to Michael A. Garrabrant.
Application Number | 20200182561 16/316478 |
Document ID | / |
Family ID | 60952173 |
Filed Date | 2020-06-11 |
United States Patent
Application |
20200182561 |
Kind Code |
A1 |
Garrabrant; Michael A. |
June 11, 2020 |
CORRUGATED TUBE-IN-TUBE HEAT EXCHANGERS
Abstract
A heat exchanger has an outer tube and an inner tube extending
through a lumen of the outer tube. The outer tube has a spiral
corrugation and an inner surface of the corrugation contacts or is
in proximity to an exterior surface of the inner tube to define a
spiral annular channel in an annular space between the inner tube
and outer tube.
Inventors: |
Garrabrant; Michael A.;
(Unicol, TN) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Stone Mountain Technologies, Inc. |
Johnson City |
TN |
US |
|
|
Assignee: |
Stone Mountain Technologies,
Inc.
Johnson City
TN
|
Family ID: |
60952173 |
Appl. No.: |
16/316478 |
Filed: |
July 7, 2017 |
PCT Filed: |
July 7, 2017 |
PCT NO: |
PCT/US2017/041041 |
371 Date: |
January 9, 2019 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
62361580 |
Jul 13, 2016 |
|
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F28F 1/426 20130101;
F28F 2210/06 20130101; F28F 1/08 20130101; F28F 1/42 20130101; F28D
7/10 20130101; F28D 7/106 20130101; F28F 1/06 20130101 |
International
Class: |
F28F 1/08 20060101
F28F001/08; F28D 7/10 20060101 F28D007/10 |
Goverment Interests
GOVERNMENT LICENSE RIGHTS
[0001] This invention was made with government support under Grant
DE-EE0006116 awarded by the Department of Energy. The government
has certain rights in the invention.
Claims
1. A tube-in-tube heat exchanger comprising an outer tube and an
inner tube extending through a lumen of the outer tube, wherein the
outer tube has at least one spiral corrugation and an inner surface
of the corrugation contacts or is in proximity to an exterior
surface of the inner tube to define a spiral annular channel in an
annular space between the inner tube and outer tube.
2. The tube-in-tube heat exchanger of claim 1, wherein the inner
tube and outer tube are concentrically arranged.
3. The tube-in-tube heat exchanger of claim 1, wherein the inner
tube has at least one spiral corrugation.
4. The tube-in-tube heat exchanger of claim 1, wherein the inner
tube is a smooth tube.
5. The tube-in-tube heat exchanger of claim 1, wherein the outer
tube has more than one spiral corrugation.
6. The tube-in-tube heat exchanger of claim 1, wherein a pitch of
the spiral corrugation is from about 0.25 to about 5.
7. The tube-in-tube heat exchanger of claim 1, wherein a height of
the annular channel is about 0.01 inches to about 0.5 inches.
8. The tube-in-tube heat exchanger of claim 1, wherein the inner
surface of the corrugation is within 0.02 inches or less of the
exterior surface of the inner tube.
9. The tube-in-tube heat exchanger of claim 3, wherein the
corrugation of the inner tube is in a direction opposite to the
corrugation of the outer tube.
10. The tube-in-tube heat exchanger of claim 1, wherein the inner
tube contains an insert that increases turbulence of a fluid
flowing inside the inner tube.
11. The tube-in-tube heat exchanger of claim 1, wherein the inner
and outer tubes are made from a metal.
12. The tube-in-tube heat exchanger of claim 1, wherein the inner
and outer tubes are made from a plastic or elastomer.
13. The tube-in-tube heat exchanger of claim 1, wherein one of the
tubes is made from metal and the other from plastic or
elastomer.
14. The tube-in-tube heat exchanger of claim 1, wherein the heat
exchanger is coiled.
15. A refrigerant sub-cooler (RSC) for a heat pump comprising the
heat exchanger of claim 1.
16. A solution heat exchanger (SHX) for an absorption heat pump
comprising the heat exchanger of claim 1.
17. A heat pump comprising the tube-in-tube heat exchanger of claim
1, wherein the heat pump is configured to provide a ratio of a
pressure loss inside the tube (DELTAP_t) to a pressure loss in the
annular channel (DELTAP_a) of 0.9 or less.
18. A heat pump comprising the tube-in-tube heat exchanger of claim
1, wherein a fluid inside the inner tube flows in the same
direction as a fluid flowing inside the annular channel.
19. A heat pump comprising the tube-in-tube heat exchanger of claim
1, wherein a fluid inside the inner tube flows in the opposite
direction as a fluid flowing inside the annular channel.
Description
TECHNICAL FIELD
[0002] This disclosure relates to tube-in-tube heat exchangers,
particularly, tube-in-tube heat exchangers having spiral
corrugations.
BACKGROUND
[0003] Tube-in-tube heat exchangers are used to transfer heat from
one fluid to another. Tube-in-tube heat exchangers usually consist
of two tubes (i.e., an inner tube and an outer tube), in which the
outermost diameter of the inner tube is smaller than the innermost
diameter of the outer tube so that the inner tube fits completely
inside the outer tube. A first fluid flows inside the inner tube,
while a second fluid flows in the annulus formed between the
exterior surface of the inner tube and the inner surface of the
outer tube. The two fluids flow and exchange heat conducted through
the wall of the inner tube, without direct contact between the two
fluids. Fluids can be in any flowable form (e.g., liquid, gas,
two-phase, with and without entrained solids and the like), and the
fluids may flow through a heat exchanger in a co-flow (i.e., same
direction) or counter-flow (i.e., opposite direction).
[0004] There are several variables that impact the rate of heat
transfer and the resulting pressure loss. For a simple "smooth"
tube-in-tube heat exchanger, the inside diameter (Di_i) of the
inner tube establishes the mass flux, velocity, convective heat
transfer coefficient, and pressure loss per unit length for the
first fluid flowing inside the inner tube. The spatial difference
between the inside diameter of the outer tube (Do_i) and the
outside diameter of the inner tube (Di_o) defines an annular gap in
which the second fluid flows. The resulting hydraulic diameter
(Do_i-Di_o) establishes the mass flux, velocity, convective heat
transfer coefficient, and pressure loss per unit length for the
second fluid flowing in the annulus.
[0005] The basic "smooth" or "straight-sided" tube-in-tube geometry
often has limitations that prevent an optimum heat exchanger
design. For example, standard tubes are only available in a limited
number of diameter and wall thickness combinations, limiting the
available options for the key geometric variables (Di_o and Di_i).
Often, the characteristics (type of fluid, thermophysical
properties and the like) of the two fluids are quite different and
the respective flow rates and allowable pressure loss may differ as
well. If one fluid is a gas that has a low pressure loss
requirement, it will almost necessarily have to flow through the
inner tube that is large in diameter. Since the flow area of the
annulus (and resulting fluid velocity) is a function of the
diameter squared, the resulting annular flow area may be larger
than desired (causing low fluid velocities and low heat transfer
rates) if the other fluid is a liquid at a low flow rate. Even if
the inside diameter of the outer tube (Do_i) and the outside
diameter of the inner tube (Di_o) can be closely matched so that
annulus side velocity increases, the pressure loss penalty for a
viscous fluid flowing through a very narrow gap is often very high.
The result is the inability to arrive at a desired heat exchanger
geometry where the size of the heat exchanger is constrained by the
annulus-side fluid heat transfer rate or resulting pressure
loss.
[0006] Additionally, it is often desirable to coil the resulting
length of concentric tubes so that the heat exchanger fits within a
smaller package size. However, when two tube-in-tube smooth tubes
are coiled, the inner tube does not remain centered (concentric)
with respect to the outer tube. Instead, its center axis shifts
towards the inside of the coil until the outside diameter of the
inner tube rests against the inside diameter of the outer tube.
While this may not impact the flow, heat transfer and pressure loss
per unit length of the fluid flowing inside the inner tube, the
resulting non-concentric annular space negatively impacts the flow
and heat transfer characteristics of the fluid flowing in the
annular space. The flow in the annulus is no longer evenly
distributed around the outside diameter of the inner tube, reducing
the effective heat transfer area and increasing the length of the
heat exchanger (and resulting pressure loss for both fluids).
[0007] Additionally, the hydraulic diameter for the resulting
non-concentric annular flow area is no longer simply defined as
Do_i-Di_o. The resulting crescent-shaped annual gap that varies
along the radial axis is a function of the flow area and the wetted
perimeter (4*A/WP), which normally results in a larger hydraulic
diameter compared to a concentric annulus and, therefore, lower
heat transfer rates.
[0008] Various methods have been provided to keep the inner tube
concentric after coiling, including helically wrapping the inner
tube with a wire (e.g., US 2015/0159957) or forming dimples in the
outer tube (e.g., U.S. Pat. No. 2,259,433). The helically wrapped
wire can introduce turbulence, mixing, and a spiral flow pattern in
the annulus, which are often desirable features that can increase
the annulus-side heat transfer coefficient. However, the annulus
side pressure loss per unit length often increases dramatically and
the cost of wrapping and attaching (normally welding or brazing)
the wire to the inner tube can be very expensive. Additionally,
while dimples are effective at keeping the inner tube concentric
and in some cases slightly increasing the annulus side heat
transfer coefficient due to increased mixing, dimples are not
suitable for a heat exchanger where the annulus side constrains the
heat exchanger design due to heat transfer or pressure loss.
[0009] Flow in smooth tubes most often results in low heat transfer
performance due to the lack of turbulence, mixing, and the buildup
of a thick boundary layer on the smooth wall. Numerous methods and
geometries to enhance heat transfer performance (normally with a
corresponding pressure loss penalty) in tube-in-tube heat
exchangers have been developed. Many of those methods utilize fins
or fin-like structures located in the annular space between the two
tubes and/or inside the inner tube, or inserts that introduce
turbulence (e.g., U.S. Pat. Nos. 2,692,763, 4,286,653, 6,098,704,
US 2010/00193168, US 2015/0159957).
[0010] When a twisted or fluted tube is used as the inner tube of a
concentric tube-in-tube heat exchanger, both the tube-side and
annulus-side heat transfer rate is increased (e.g., US
2009/0159248). However, the geometry of fluted tubes (flute height,
pitch, number of flutes) is constrained by fabrication and material
property limitations. Due to such geometrical constraints, it is
often not possible to fabricate a concentric tube-in-tube heat
exchanger with an inner fluted tube that provides the desired
balance between heat transfer enhancement and pressure loss for
both fluids.
[0011] Accordingly, there remains a need for a heat exchanger that
can maintain a concentric tube-in-tube arrangement, enhance heat
transfer and minimize pressure loss for both fluids.
SUMMARY
[0012] I provide a tube-in-tube heat exchanger comprising an inner
tube and an outer tube. The outer tube has at least one spiral
corrugation and an inner surface of the corrugation is in contact
with or in proximity to the exterior surface of the inner tube to
provide a spiral annular channel in an annular space between the
inner tube and outer tube.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1A depicts an example of a tube-in-tube heat exchanger
in which the outer tube is corrugated and the inner tube is not
corrugated.
[0014] FIG. 1B depicts the cross-sectional view of the tube-in-tube
heat exchanger of FIG. 1A.
[0015] FIG. 2 depicts an alternative example of a tube-in-tube heat
exchanger in which both the outer and inner tubes are
corrugated.
[0016] FIG. 3 depicts a cross-sectional view of the annular flow
area formed between two adjacent corrugations, the inside diameter
of the outer tube, and the outside diameter of the inner tube.
DETAILED DESCRIPTION
[0017] It will be appreciated that the following description is
intended to refer to specific examples of structure selected for
illustration in the drawings and is not intended to define or limit
this disclosure other than in the appended claims.
[0018] This disclosure provides a tube-in-tube heat exchanger where
the heat transfer performance and pressure loss of two fluids
flowing through the heat exchanger can be controlled by specifying
the diameter of the inner and outer tubes and the corrugation pitch
of the outer tube. This disclosure also provides a simple and
inexpensive method of maintaining concentricity of the inner tube
if the tube-in-tube heat exchanger is bent or coiled. The resulting
heat exchanger geometry is simple and inexpensive to manufacture,
provides for a heat exchanger that can be controlled for the
desired performance, and is easily adaptable to a wide range of
applications by setting the desired diameter and pitch
variables.
[0019] Referring to the drawings, FIGS. 1A and 1B depict an
exemplary tube-in-tube heat exchanger 100 comprising an inner tube
102 and an outer tube 104. The inner tube 102 is concentrically
positioned in the lumen of the outer tube 104 and the outer tube
104 has at least one spiral corrugation 108. The corrugation 108 is
defined by a groove formed in the tube wall and extending in a
spiral pathway. A groove on the exterior surface of the tube
corresponds to a ridge on the interior surface of the tube. The
distance between corrugations 108 (pitch) is indicated by "P."
[0020] As shown in FIG. 1A, the inner surface 106 of the
corrugation 108 of the outer tube 104 contacts or is in proximity
to the exterior surface of the inner tube 102. A suitable distance
between the inner surface 106 of the corrugation 108 and the
exterior surface of the inner tube 102 for the structures to be "in
proximity" may be about 0.02 inches or less, preferably, about 0.01
inches or less or, even more preferably, about 0.005 inches or
less.
[0021] The arrangement of the inner surface 106 of the spiral
corrugation 108 of the outer tube 104 in contact with or in
proximity to the exterior surface of the inner tube 102 forms a
spiral annular channel 110 in the annular space between the inner
tube 102 and outer tube 104. Fluid in the annular channel 110 can
flow in a spiral path around the outside of the inner tube 102 in
the annular gap formed between adjacent portions of the corrugation
108, the inner surface of the outer tube 104, and the exterior
surface of the inner tube 102. The spiral flow direction of fluid
in the annular gap is depicted by the bold arrows in FIG. 1A.
[0022] FIG. 3 depicts a cross-sectional view of the annular channel
110. The pitch of the spiral corrugation 108, along with the
outside diameter of the inner tube 102 (Di_o) and the inside
diameter of the corrugated outer tube 104 (Do_i), determines the
cross-sectional area of the annular channel 110. As shown in FIG.
3, the cross-sectional shape of the annular channel 110 may
approximate a distorted rectangle or trapezoid with rounded,
tapered sides. The area of resulting annular channel 110 determines
the velocity of the fluid flowing in the annular channel 110.
[0023] Length "g" defines the height of the annular channel 110.
Length "g" is the distance between the inner surface of the outer
tube 104 and the exterior surface of the inner tube 102. A suitable
length "g" may be at least 0.01 inches. A preferred range for "g"
is about 0.01 inches to about 0.5 inches or, more preferably, about
0.01 inches to about 0.1 inches.
[0024] P* is the actual width of the flow area based on the tube
diameter and corrugation pitch (P). The perpendicular length (P*)
between the adjacent portions of the corrugation 108 is calculated
based on the corrugation pitch (P) and the outside diameter of the
inner tube 102 (Di_o):
P*=P.times.sin(.THETA.)
.THETA.=arctan(Di_o.times.3.14/P)=helix angle of corrugation
If there is more than one corrugation (i.e., more than one spiral
groove), P* is calculated based on the above equation and divided
by the number of corrugations.
[0025] The flow velocity and dimensions of the annular channel 110
are key factors determining the heat transfer performance and
pressure loss for the fluid flowing in the annular channel 110.
Therefore, by varying the outside diameter of the inner tube 102
(Di_o), the inside diameter of the outer tube 104 (Do_i), the
length between corrugations (pitch) and ensuring the inner surface
106 of the corrugation 108 is in contact with or is in proximity to
the exterior surface of the inner tube 102, the heat transfer
performance and pressure loss for the fluid flowing in the annular
space can be precisely controlled.
[0026] The corrugations 108 in the outer tube 104 also provide a
centering mechanism for the inner tube 102 if the concentric tubes
(102 and 104) are coiled so that desired heat transfer performance
is maintained.
[0027] The geometry of the inner tube 102 can be varied depending
upon the application and the desired heat exchanger performance.
The inner tube 102 can be a simple smooth tube that provides no
heat transfer enhancement, or a corrugated tube to provide heat
transfer enhancement to the fluid flowing in the inner tube 102.
FIG. 2 illustrates an example of a tube-in-tube heat exchanger 200
in which both the inner tube 202 and outer tube 204 are corrugated.
The inner surface 206 of the corrugation 208 is in contact with or
is in close proximity to the exterior surface of the inner tube
202. In FIG. 2, the corrugation direction of the outer tube 204 is
opposite that of the inner tube 202.
[0028] If a corrugated tube is used for both the inner tube 102 and
the outer tube 104, the corrugation pitch (P) may be the same or
different for the tubes. A suitable pitch for the corrugations of
either the inner tube 102 or the outer tube 104 may be from about
0.25 to about 5, more preferably, from about 0.3 to about 3. The
corrugation depth may be the same or different for the two tubes.
The direction of spiral may be the same or different (one clockwise
and the other counter-clockwise for example).
[0029] Other forms of heat transfer enhancement can be used for the
fluid flowing inside the inner tube 102, totally independent of the
annular flow geometry in the annular channel 110, including
internally formed fins, ridges, inserts or the like that promote
turbulence and heat transfer enhancement.
[0030] The number of corrugation grooves in the outer tube or inner
tube (if corrugated) can be one or more than one. If there are two
corrugation grooves, the annular flow will be divided into two
adjacent spiral flow paths defined by the gap between outer tube
and inner tube and the distance between corrugations. If there are
three corrugation grooves, the annular flow will be divided into
three flow paths, and so on. Preferably, the number of corrugation
grooves is between 1 and 4.
[0031] The material of the tubes is not particularly limited and
can be selected based on the desired application of the heat
exchanger. Suitable materials include metal, plastic, elastomer and
the like. The tubes of a heat exchanger may be made from the same
material or from different materials.
[0032] The tube-in-tube heat exchangers are especially useful for
heat exchangers in which the two fluids flowing through the heat
exchanger are of different forms (liquid or gas, for example), have
different flow rates, have different pressure loss specifications,
and/or a phase change is occurring for one of the fluids
(evaporation or condensation). This is due to the fact that the
heat transfer rates can be balanced between the two fluids by
choosing the tube diameters and corrugation pitch. Preferably, the
fluid that must have a low pressure loss through the heat exchanger
is selected to flow inside the inner tube and the fluid that can
withstand a higher pressure loss through the heat exchanger is
selected to flow through the annular channel, where the velocity
can be controlled by the chosen geometry to be higher. Preferably,
the pressure loss inside the tube (DELTAP_t) is less than the
pressure loss in the annulus (DELTAP_a). In preferred examples, a
ratio of (DELTAP_t):(DELTAP_a) is about 0.9 or less, more
preferably, about 0.8 or less or, even more preferably, about 0.5
or less.
[0033] A heat exchanger application where the tube-in-tube heat
exchangers are especially useful is a refrigerant sub-cooler (RSC)
for a heat pump. In an RSC, one fluid is the cold, low-pressure
refrigerant vapor exiting the evaporator which may still contain
some amount of unevaporated liquid. The other fluid is hot
high-pressure refrigerant liquid exiting the condenser. In the RSC,
the hot high-pressure liquid is cooled to a temperature below its
saturation temperature by the cold low pressure refrigerant 2-phase
mixture.
[0034] The heat transfer coefficient for the low pressure 2-phase
mixture will be very high without enhancement due to evaporation of
the remaining liquid as it passes though the RSC. However, it is
desirable to keep the pressure loss of the low pressure 2-phase
refrigerant as low as possible to ensure maximum efficiency of the
heat pump cycle. Therefore, it is desirable for the low-pressure
refrigerant flow to be inside the inner tube 102 of sufficient size
so the velocity is low and pressure loss is minimized. Since the
allowable pressure loss of the high-pressure refrigerant liquid
flowing in the annulus is much higher than the low-pressure 2-phase
refrigerant flowing inside the inner tube, the corrugated
tube-in-tube geometry allows the designer to take advantage of this
design condition and a shorter (less expensive) heat exchanger
results.
[0035] In the RSC, the heat transfer coefficient for the
high-pressure refrigerant liquid will be much lower given that no
phase-change process occurs. Since the fluid is a liquid and,
therefore, of much higher density than the low pressure vapor, it
can flow through a much smaller flow area without a large pressure
loss penalty. Also, since the high-pressure refrigerant liquid will
be expanded to the low-side pressure before entering the evaporator
after exiting the RSC, the high-pressure liquid refrigerant can
have a high-pressure loss through the RSC without impacting the
efficiency of the heat pump.
[0036] Therefore, the diameter of the inner tube can be selected to
provide the desired (low) pressure loss for the low-pressure
refrigerant. The outside diameter of the outer tube and the pitch
of the corrugations in the outer tube are then selected so that the
velocity of the high-pressure refrigerant liquid is high and the
heat transfer performance of the high-pressure liquid flowing
spirally in the annulus closely matches that of the low pressure
2-phase fluid. The high velocity of the high-pressure refrigerant
liquid can cause a high-pressure loss through the RSC but, in this
application, that is acceptable.
[0037] Another heat exchanger application where the tube-in-tube
heat exchangers are especially useful is the solution heat
exchanger (SHX) in an absorption heat pump. In the SHX, thermal
energy is transferred and recuperated from the hot "weak solution
(WS)" exiting the Desorber (solution with low refrigerant content)
to the cool "strong solution (SS)" entering the Desorber (solution
with high refrigerant content). Although both fluids are at the
high-side pressure, the SS has a higher flow rate than the WS and
it is preferred to have a low pressure loss through the SHX so that
additional back pressure is not put on the solution pump
(increasing its required power and reducing reliability). Since the
WS will be expanded to the low-side pressure after exiting the SHX,
a high pressure loss for the WS in the SHX is acceptable.
[0038] Therefore, the diameter of the inner tube can be selected to
provide the desired pressure loss for the SS flowing inside the
inner tube. Since the SS is a liquid, use of a corrugated inner
tube or an insert inside the inner to tube to induce turbulence is
desirable. The outside diameter of the outer tube and the pitch of
the corrugations in the outer tube are then selected so that the
heat transfer performance of the WS flowing in the annulus matches
or is higher than the SS flowing inside the enhanced inner tube.
The resulting high velocity and, therefore, high-pressure loss of
the WS is acceptable because a high-pressure loss for the WS
flowing through the SHX does not impact the efficiency of the heat
pump.
[0039] Many other heat exchanger applications where the two fluids
are mismatched in flow rate, phase or allowable pressure loss can
benefit from the heat exchanger.
EXAMPLES
Examples 1-17
[0040] Solution heat exchangers (SHX) for a nominal 80,000 Btu/hr
gas absorption heat pump (GAHP) were designed (sized) using a
corrugated inner tube inserted into a corrugated outer tube where
the inner surface of the corrugations of the outer tube were in
contact with or in proximity to the exterior surface of the inner
tube. The strong solution was selected to flow inside the inner
tube and the weak solution was selected to flow inside the annular
channel.
[0041] For Examples 1-17, the outside diameter and wall thickness
of both the inner and outer tubes was varied, along with the pitch
of the corrugations in the outer tube. The calculated heat
exchanger lengths ranged from 18 to 60 feet. The geometries of the
heat exchangers of Examples 1-17 are shown in Table 1.
[0042] By selecting different tube diameter and wall thickness
combinations, different annular gap (g) dimensions can be obtained
and these gaps are short enough to allow high velocities (and
therefore high heat transfer coefficients) for the weak solution
flowing in the annulus. The corrugation pitch was also varied to
create smaller or larger flow areas (and therefore velocities) in
the annulus.
[0043] The pressure loss inside the tube (DELTAP_t), the pressure
loss in the annulus (DELTAP_a), and overall heat transfer
coefficient (U_o) for each example are shown in Table 1.
TABLE-US-00001 TABLE 1 Inner Outer diameter Outer diameter
Thickness of outer diameter of Thickness of outer of outer
tube(Do_i) inner tube of inner "g" Example# tube (in) tube (in)
(Di_o) (in) tube (in) (in) Pitch 1 0.5 0.028 0.444 0.375 0.028
0.035 1.00 2 0.5 0.035 0.430 0.375 0.035 0.028 1.00 3 0.5 0.049
0.402 0.375 0.049 0.014 1.00 4 0.5 0.035 0.430 0.375 0.049 0.028
3.00 5 0.5 0.035 0.430 0.375 0.049 0.028 2.00 6 0.5 0.035 0.430
0.375 0.049 0.028 1.50 7 0.5 0.035 0.430 0.375 0.049 0.028 1.25 8
0.5 0.035 0.430 0.375 0.049 0.028 0.75 9 0.625 0.049 0.527 0.375
0.049 0.076 0.50 10 0.625 0.049 0.527 0.375 0.049 0.076 0.33 11
0.625 0.065 0.495 0.375 0.049 0.060 0.33 12 0.625 0.035 0.555 0.5
0.049 0.028 1.00 13 0.625 0.035 0.555 0.5 0.065 0.028 1.00 14 0.625
0.035 0.555 0.5 0.065 0.028 1.50 15 0.625 0.035 0.555 0.5 0.065
0.028 0.75 16 0.75 0.049 0.652 0.5 0.049 0.076 1.00 17 0.75 0.065
0.620 0.5 0.049 0.060 1.00 U_o alpha_a alpha_t Length of DELTAP_t
DELTAP_a [Btu/hr- [Btu/hr- [Btu/hr- Example# tube (ft) (psi) (psi)
ft{circumflex over ( )}2-F] ft{circumflex over ( )}2-F]
ft{circumflex over ( )}2-F] 1 23 1.2 15.9 616 2130 1104 2 21 1.4
29.1 659 2667 1197 3 18 1.9 207.7 767 5413 1424 4 60 6.3 8.0 235
318 1424 5 34 3.6 13.5 412 766 1424 6 24 2.5 16.2 587 1715 1424 7
22 2.3 20.5 637 2220 1424 8 20 2.1 45.1 703 3295 1424 9 24 2.5 5.9
581 1659 1424 10 22 2.3 11.8 644 2310 1424 11 21 2.2 21.4 683 2902
1424 12 24 0.4 32.2 446 2667 730 13 23 0.6 31.6 456 2667 847 14 25
0.7 17.2 416 1715 847 15 22 0.6 50.5 471 3295 847 16 31 0.5 2.1 346
977 730 17 28 0.5 3.9 374 1233 730
Comparative Examples 1-8
[0044] Solution heat exchangers (SHX) for a nominal 80,000 Btu/hr
gas absorption heat pump (GAHP) were designed (sized) using an
inner fluted (or twisted) tube inserted into a smooth outer tube,
where the inside diameter of the outer tube is in contact with or
very close proximity to the outside diameter of the inner fluted
tube (tips of the flutes). The dimensions of the heat exchangers of
Comparative Examples 1-8 are shown in Table 2. These tubes were
selected based on having the shortest flute heights (e) believed
possible based on the data base of manufacturable fluted tube
geometries published in "A Manual for Heat Exchanger Design Using
Spirally Fluted Tubes," under Gas Research Institute contract
#5092-243-2357. The shorter flute heights (g) provide the highest
velocity and heat transfer coefficients for the weak solution
flowing in the annulus. The heat exchanger lengths ranged from 142
to 252 feet as shown in Table 2, representing a significant
increase in heat exchanger size compared to the corrugated tubes of
Examples 1-17.
[0045] In the SHX, hot weak solution (a liquid low in refrigerant
concentration) transfers heat to a strong solution (a liquid of
high refrigerant concentration). The flow rate of the strong
solution is higher than the weak solution. The allowable pressure
loss of the strong solution was kept very low (to prevent
back-pressure on the pump), while the allowable pressure loss of
the weak solution can be much higher. For these reasons, the strong
solution was designated to flow inside the inner tube and the weak
solution flows in the annulus between the inner fluted tube and the
outer plain tube.
[0046] The pressure loss inside the tube (DELTAP_t), the pressure
loss in the annulus (DELTAP_a), and overall heat transfer
coefficient (U_o) were calculated. Results of SHX designs using the
eight selected fluted tube geometries are shown in Table 2.
[0047] For all Comparative Examples 1-8, the heat exchanger size is
limited by the heat transfer coefficient in the annulus (alpha_a),
which is much lower than inside the tube (alpha_t). This result can
be seen by the overall heat transfer coefficient (U_o) which is
very close to alpha_a. Also, for most of Comparative Examples 1-8,
the pressure loss inside the tube (DELTAP_t) is higher than the
pressure loss in the annulus (DELTAP_a), which is the opposite of
what is desired.
TABLE-US-00002 TABLE 2 Inner Outer diameter Outer diameter
Thickness of outer diameter of Comp. of outer of outer tube(Do_i)
inner tube "g" Number Example# tube (in) tube (in) (Di_o) (in) (in)
of starts Pitch 1 0.740 0.020 0.700 0.402 0.149 3 0.522 2 0.690
0.020 0.650 0.445 0.103 4 0.545 3 0.655 0.020 0.615 0.420 0.098 4
0.324 4 0.636 0.020 0.596 0.366 0.115 4 0.261 5 0.560 0.016 0.528
0.333 0.098 3 0.522 6 0.505 0.016 0.473 0.318 0.078 4 0.24 7 0.785
0.028 0.729 0.589 0.070 4 0.6 8 0.780 0.020 0.740 0.570 0.085 5
0.245 U_o alpha_a alpha_t Comp. Length of DELTAP_t DELTAP_a
[Btu/hr- [Btu/hr- [Btu/hr- Example# tube (ft) (psi) (psi)
ft{circumflex over ( )}2-F] ft{circumflex over ( )}2-F]
ft{circumflex over ( )}2-F] 1 223 3.3 0.2 44 47 814 2 214 2.7 0.4
44 47 729 3 190 3 0.4 52 55 1017 4 194 6.1 0.3 57 60 1215 5 252 9.4
0.7 49 51 1119 6 208 12.5 1.1 63 66 1657 7 171 0.4 0.5 43 47 599 8
142 0.7 0.3 54 58 802
[0048] For the corrugated arrangements of Examples 1-17, the
annulus side heat transfer coefficients (alpha_a) and pressure
losses (DELTAP_a) are much higher than for the fluted tube designs
of Comparative Examples 1-8, resulting in a significantly higher
overall heat transfer coefficient (U_o) and much shorter heat
exchanger lengths. Also, the pressure loss in the annulus
(DELTAP_a) is higher than the inside tube (DELTAP_t). Since the
allowable pressure loss of the weak solution flowing in the annulus
is much higher than the strong solution in the tube, the corrugated
tube-in-tube geometry allows the designer to take advantage of this
design condition and a shorter (less expensive) heat exchanger
results.
Examples 18-19
[0049] Two heat exchangers were fabricated and tested in a GAHP
prototype using the tube and heat exchanger dimensions shown in
Example 12. The heat exchanger of Example 18 was 20 foot long (4
foot shorter than design) and provided an effectiveness of 0.92
(0.97 design target). The heat exchanger of Example 19 was 30 foot
long and provided an effectiveness of 0.98.
[0050] Although the apparatus and methods have been described in
connection with specific forms thereof, it will be appreciated that
a wide variety of equivalents may be substituted for the specified
elements described herein without departing from the spirit and
scope of this disclosure as described in the appended claims.
* * * * *