Title:
Multiple concentric cylindrical co-coiled heat exchanger
Kind Code:
A1


Abstract:
A compact shell and coil heat exchanger is disclosed that accommodates widely differing volumetric flowrates between the two fluids undergoing heat exchange. Multiple co-coiled helical coils of tubing are concentrically arranged, and coil spacers are provided which maintain the tubes in overall staggered alignment, as illustrated in FIG. 1. Uniformly high transfer coefficients are maintained throughout the bundle of coils via means for ensuring that the tube-side flow through the tubes of each coil, and the shell-side flow across each coil, are kept proportional to the tube surface area of each coil.



Inventors:
Erickson, Donald Charles (Annapolis, MD, US)
Application Number:
12/587373
Publication Date:
04/22/2010
Filing Date:
10/06/2009
Primary Class:
Other Classes:
165/163
International Classes:
F28D7/12; F28D7/02
View Patent Images:
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Primary Examiner:
TEITELBAUM, DAVID J
Attorney, Agent or Firm:
Donald C. Erickson (Annapolis, MD, US)
Claims:
I claim:

1. A heat exchanger comprised of a) Multiple concentric helical coils of tubing, wherein each coil is coiled in the same direction; b) A pressure vessel containing said coils; c) At least one spacer for each coil, which maintains approximately equal tube gaps within the successive loops of the coil, and approximately equal coil-to-coil gaps between it and an adjoining coil; d) Connections between adjoining spacers which maintain the spacers in alignment such that the tubes of adjoining coils are in staggered alignment; and e) A tube header for each end of the collection of coils, which connects the tubes at that end to an external fluid conduit.

2. The apparatus according to claim additionally comprising a core blocker that prevents shell-side fluid from bypassing the bundle of multiple coils.

3. The apparatus according to claim 2 wherein said tubes are are circular cross section tubes with diameters in the approximate range of 3/160 to ¾ inches.

4. The apparatus according to claim 2 wherein said coil-to-coil gap is in the approximate range of 0.1 D to 0.8 D, and the tube gap is in the approximate range of 0.3 D to 1 D, where D is the diameter of the tubing.

5. The apparatus according to claim 4 wherein said tubes include at least one of grooving and corrugations.

6. The apparatus according to claim 3 adapted to be a refrigerant heat exchanger in an absorption refrigeration cycle, and including a bottom entry in the shell for a cold refrigerant vapor, and a warm refrigerant liquid connection to the top tube header, whereby refrigerant vapor flows axially upward through the shell, and refrigerant liquid flows countercurrently downward through the tube bundle.

7. The apparatus according to claim 3 adapted to be a heat recovery vapor generator in an absorption cycle for producing at least one of refrigeration, heat pumping, and power, and including a bottom entry in the shell for a hot exhaust gas, a preheated absorbent liquid solution connection to the top tube header, and an outlet for partially desorbed solution from the bottom tube header, whereby exhaust gas flows axially upward through the shell, and desorbing liquid flows countercurrently downward through the tube bundle.

8. The apparatus according to claim 3 adapted to be an evaporator for chilling a gas such as air in an absorption refrigeration cycle, and including a top entry in the shell for the gas to be chilled, and a pre-cooled refrigerant liquid connection to the bottom tube header, whereby the gas being chilled flows axially downward through the shell, and evaporating refrigerant liquid flows countercurrently upward through the tube bundle.

9. The apparatus according to claim 3, adapted to be a component of a closed ammonia-water absorption cycle, with ammonia-containing liquid supplied to one of the tube headers, and low pressure vapor supplied to the shell at the other end of the tube bundle.

10. A heat exchanger comprising: a) A bundle of multiple concentric coils of helically coiled tubing, wherein at least some of the coils have more than one tubing start, and wherein all the coils are wound in the same direction b) A pressure containment for said bundle of coils, which admits a first fluid to one end of said bundle, c) A tube header for each end of said bundle, which admits a second fluid into and out of said coils.

11. The apparatus according to claim 10 additionally comprised of a central core blocker that prevents said first fluid from bypassing said bundle.

12. The apparatus according to claim 11 additionally comprised of spacers for each coil which maintain approximately equal tube gaps and approximately equal coil gaps, and connectors for said spacers which maintain most of the tubes of adjacent coils in staggered alignment.

13. The apparatus according to claim 12 additionally comprising a second tube header at each end of said bundle, which connects a third fluid to a subset of said coils.

14. The apparatus according to claim 12 wherein the diameter of said tubes is in the range of 3/16 to ⅝ inches, and wherein said tube gaps are in the range of 0.1D to 1D, and the coil gaps are in the range of D/20 to ¾ D.

15. The apparatus according to claim 12 additionally comprised of spacer strips on said blocker which maintain a gap between the blocker and the innermost coil of approximately one half the coil-to-coil gap, plus spacers that maintain the gap between the outermost coil and the containment shell at approximately one half the coil-to-coil gap.

16. The apparatus according to claim 12 wherein the tube flow through and shell flow around the tubes of each coil are maintained approximately proportional to the tube surface area of the coil by at least one of: a) Placing flow-reducing inserts in the shorter tubes; b) Using smaller diameter tubes in the smaller diameter coils; and c) Increasing the number of tube starts in the larger coils proportional to the coil diameter.

17. A shell and coil heat exchanger comprised of multiple co-coiled concentric coils of tubing with multiple tube starts in each coil, adapted for use in an ammonia-water absorption cycle apparatus as at least one of the refrigerant heat exchanger, the heat recovery vapor generator, and the evaporator.

18. The apparatus according to claim 17 additionally comprised of at least one spacer for each coil that maintains the tubes of that coil in staggered alignment with the tubes of an adjacent coil.

19. A shell and coil heat exchanger with multiple co-coiled concentric coils of tubing characterized in that each coil of tubing is adjusted to have tube-side flow through the tubing and shell-side flow around the tubing which is proportional to the tube surface area of that coil, the same proportionality holding for all the coils, by at least one of: a) Placing flow-reducing inserts in the shorter tubes; b) Using smaller diameter tubes in the smaller diameter coils; and c) Increasing the number of starts in the larger coils proportional to the coil diameter.

20. The apparatus according to claim 19 additionally comprised of spacers for each coil which are interconnected so as to maintain overall staggered alignment of the tubes.

Description:

CROSS REFERENCE TO RELATED APPLICATIONS

Not Applicable

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH AND DEVELOPMENT

Not Applicable

BACKGROUND OF THE INVENTION

Shell and coil heat exchangers have found wide-ranging use in commerce and industry. Whereas various shell geometries and coil geometries are found in the prior art, the ones of interest here are cylindrical shells which contain helically coiled tubing, where the tube coils are also cylindrical in shape. One prior art example of a single cylindrical coil contained in a cylindrical shell is found in U.S. Pat. No. 6,044,837. Note the central core flow blocker, which forces all shell-side flow through the helical coil region. Note also the wide gap between adjacent loops of the tube coil (the “tube gap”). That gap provides the flow space for the shell side fluid, which follows a helical path countercurrent to the tube fluid flowpath.

Of particular interest for this disclosure are the cylindrical shell helical tube coil geometries that contain multiple co-coiled concentric helical coils of tubing in a single shell. Examples of that are found in the following U.S. Pat. Nos. 2,888,251; 4,106,309; 4,193,268; 4,556,103; 4,865,124; 5,088,192; 5,228,505; 5,379,832; 5,490,393; 6,560,966; and 7,096,664. All of these geometries have the central core blocker, and have multiple concentric helical coils of tubing which are co-coiled. “Co-coiled” means that all of the tubes are coiled in the same direction. Also, in the above references each individual helical coil has a single start of tubing.

Note that when there are multiple concentric helical single start tube coils, all with the same tube gap, the length of tube in the outer coils is inherently longer than the length of tube in the inner coils, because of their greater circumference. This has been found to cause a serious flow imbalance. On the tube side, more of the tube flow goes through the inner tube, due to its shorter length and lower pressure drop. However on the shell side, more of the flow goes through the portion of shell containing the outer tube, since that represents a larger portion of the total shell-side flow area. The tube gaps could be increased for the shorter inner tubes, thus increasing shell flow across those tubes, but that would not solve the problem, as the tubes would get even shorter, so there would be even less transfer area and larger tube side flow.

Five of the above disclosures provide a corrective measure for the above imbalance. In the '103, '124, '832, '505, and '192 disclosures, halfway up the tube bundle, the outermost coil and innermost coil trade places. This roughly equalizes the total length of the tubes, and also the shell-side flow area for each. Note however that this requires complicated re-circuiting in the middle of the bundle(s) of coils.

Other disclosures of concentric helical tube coil heat exchangers of interest include U.S. Pat. Nos. 1,526,320; 6,095,240; and 6,568,467. Note that each of these discloses multiple concentric cylindrical shells, with each shell generally only containing a single tube coil.

There are well-known cost and performance advantages associated with making heat exchangers more compact. However the prior art shell and coil exchangers with multiple co-coiled concentric helical coils encounter a problem when striving for compactness. Note that the coil bundle geometry can be characterized by two gaps—the gap between adjacent coils (assuming all are approximately equal), and the gap between adjacent tubes in a single coil (again assuming all are approximately equal). As the tube gap decreases to zero, as illustrated in the '393 patent, the shell side becomes divided into distinct annuli. Thus the coil-to-coil gaps define the only flow area available on the shell side. If the coil gaps also go to zero, and the tubes are arranged in an in-line configuration, i.e. each tube lined up with its neighboring tube in the adjacent coil(s), very little flow is possible. Conversely, when the coil gaps are near zero, with sizable tube gaps, as in the '268 patent, the shell-side flow can be forced into a helical flow path rather than axial, greatly increasing shell-side pressure drop. In summary, with prior art multiple co-coiled in-line configurations, the gaps cannot be made very small, and hence the bundle cannot be very compact.

These problems were countered, and the shell and coil geometry was made very compact, by adopting a counter-coiled geometry. Each helical coil is coiled the opposite direction as its neighboring coil(s). U.S. Pat. No. 6,679,083 is one example of this. Note that both the tube gap and the coil gap can be reduced all the way to zero, and there is still appreciable shell-side flow area available. When both those gaps are zero, the geometric packing density is pi divided by four. That is, 78.5% of the shell volume is occupied by tubing, and the remaining 21.5% is shell flow area. Commercial counter-coiled designs achieve tube packing densities close to 70%. That contrasts with the disclosures of co-coiled bundles with in-line tube alignment, which can at best only achieve about 35% tube packing density at the same shell-side pressure drop.

Unfortunately the counter-coiled configuration does not satisfy all heat exchange requirements. In particular, there are heat exchange applications wherein there is a great disparity between the flow area required on one side of the exchanger versus that on the other. For example, one side could be a vapor or gas at relatively low pressure, whereas the other side is a liquid. In countercurrent heat exchange it is frequently desired to have roughly equal temperature changes of the fluid on each side of the exchanger. Then countercurrent heat exchange between fluids of greatly differing densities requires greatly differing flow areas on the two sides of the heat exchange surface. That is necessary to maintain reasonable velocities, pressure drops, and transfer coefficients on both sides. For the shell and coil design, that can be achieved by placing the less dense fluid on the shell side, and providing a very low tube count on the tube side, i.e. making each tube quite long. When this (low tube count) is done with the counter-coiled configuration, the tube crossings become quite sparse, i.e. widely separated. The tubes in adjoining coils are lined up for a distance, and then staggered for a distance, with no readily achievable regular pattern. Since there is much higher flow resistance at the in-line regions compared to the staggered regions, and the various in-line regions can randomly accumulate more in some areas than in others, serious flow maldistribution can result on the shell side. What is needed, and one major objective of this invention, is to realize a shell and coil heat exchanger geometry that, in addition to the inherent advantages of that class of heat exchanger (can be all-welded, high pressure capability, wide materials selection, very accommodating to thermal expansion, ease of fabrication), can also accommodate widely differing flow areas on each side of the exchanger, without sacrificing compactness or performance.

BRIEF SUMMARY OF THE INVENTION

The above and other useful objects are achieved by modifying the multiple concentric co-coiled tubing heat exchanger to have overall staggered tube alignment coupled with small (less than one tube diameter) gaps, and by incorporating features that cause the flow through the tubes of each coil and around those tubes to be approximately proportional to the tube surface area of that coil, with the same approximate proportionality for all the coils.

BRIEF DESCRIPTION OF THE SEVERAL VIEW OF THE DRAWINGS

FIG. 1 is a cross sectional view of a first embodiment of the disclosed heat exchanger.

FIG. 2a illustrates a co-coiled tube bundle of eight concentric helical coils of tubing plus two rows of spacers and the core blocker. FIG. 2b shows the same bundle with three rows of spacers.

FIG. 3 is a cutaway view showing just the spacers plus core blocker.

FIG. 4 is a cross sectional view of the spacers on one side only, showing how they maintain the staggered alignment of adjacent tubes, in addition to maintaining the tube gaps and coil gaps.

FIG. 5 is an end view of a single spacer.

FIG. 6 is a side view of a central portion of one spacer, showing the shape of the cutouts that accept and hold the tubes.

FIG. 7 is an example flowsheet of an absorption cycle for producing chilling and/or power from exhaust heat, which can beneficially use the disclosed improved heat exchanger, as any of the refrigerant heat exchanger, the heat recovery vapor generator, and the evaporator.

FIG. 8 is a multi-bundle configuration in a single shell, where each tube bundle has multiple co-coiled concentric helical staggered coils, and the shell fluid flows sequentially through the bundles.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 illustrates the basic overall heat exchanger configuration disclosed: cylindrical shell 1 with shell-side inlet 2 and outlet 3; multiple co-coiled concentric helical coils of tubing 4, 5, 6, and 7; core blocker 8; tube headers 9 and 10 for the tubes at each end of the bundle of tube coils; and the tube header connectors 11 and 12. The critical features that enable the compactness are the close proximity or spacing of adjacent tubes, plus their staggered alignment. That close spacing and alignment is established and maintained by spacers, which are shown in subsequent figures. The illustrated shell-side upflow arrangement is the preferred configuration when it is used as a refrigerant heat exchanger.

Whereas FIG. 1 illustrates gathering all the tubes at each end of the bundle to a single tubesheet and header, the artisan will recognize that it is possible to divide the tubes into two or more groups, and provide a separate pair of tubesheets and headers for each group. That way there can be more than one liquid on the tube side, kept hermetically separate one from the other.

FIG. 2a illustrates a bundle of eight co-coiled helical concentric coils plus two rows of spacers 13 and 14, plus a central blocker. Each of the eight coils has a single tube start. Each of the two spacers for each coil are located at 180 degree opposed locations. All spacers on each side of the bundle are lined up so they can be connected one to another e.g. at the projecting ends above and below the tube bundle, thus maintaining the staggered alignment of the tubes. FIG. 2b has three rows of spacers.

FIG. 3 is a cutaway view of only the core blocker and the 16 spacers.

FIG. 4 shows in cross-section how the eight spacers on one side of the bundle are aligned so as to maintain the staggered tube alignment.

FIG. 5 is an end view of a single spacer, and FIG. 6 is a cutaway side view of a central section of two spacers, showing the cutout holes that accept and position the tubes. Note that the hole is shaped to encircle somewhat more than 180 degrees of the tube circumference, preferably between about 190 and 240 degrees. This keeps the spacer firmly adhered to the tube once it has been placed. When the FIG. 6 spacers are scaled for 5/16 inch diameter tubes (0.3125 inches), and the coil gap 61 is 0.0607 inches (0.194D), and the tube gap 62 is 0.2628 inches (0.841D) where D is the tube diameter. The included angle of each hole in this particular spacer is 233 degrees.

With a single tube start in each coil, as illustrated in FIGS. 1 and 2, the tube in each coil gets longer in proportion to the coil diameter. If not compensated for, this can be very detrimental to overall performance. Consider two coils, where one is twice the diameter of the other. The outer coil tube will be twice as long as the inner coil tube. Thus the inner tube will have appreciably less flow resistance, and hence get appreciably more tube-side flow than the outer one. On the other hand, the outer coil has twice as much shell-side flow area as the inner coil, and hence the inner coil gets much less shell-side flow than does the outer coil. It is desirable to maintain roughly the same ratio between shell-side flow, tube-side flow, and tube surface area in all parts of the exchanger, to maximize thermal performance. That is obviously not possible with a large number of same-tube diameter single start coils per se.

As mentioned above, the prior art discloses one means of overcoming this imbalance—interchanging outer and inner coils halfway along the length of the coil. It has now been discovered that there are three simpler and more preferable ways to accomplish the same objective. First, inserts can be placed inside the shorter tubes, such as wires of varying diameter, or twisted ribbons. These are preferably inserted before the tube is coiled. The insert partially blocks the flow, to the point where the tube-side flow decreases to approximately the same proportion of total tube length as that proportion for the outer tube without inserts. For the above example, with an outer tube without inserts and an inner tube with half the coil diameter, reducing the inner tubes flow to half that of the outer tube requires blocking almost three quarters of the inner tube flow area (accounting for increased friction factor). For example, for 5/16 inch (outside) diameter tubes with 0.032 inch wall thickness, the ID is 0.2485 inches, and the blocking wire or tube would need to be about 0.2 inch diameter.

A second method to maintain proportionality between the flow through the tube(s) of a single coil, shell flow around those tubes, and surface area of those tubes, is to decrease the diameter of the tubes in the smaller diameter coils, optionally in conjunction with changing their spacing (i.e. their gaps).

The third, and most preferred method, is to periodically vary the number of tube starts in the coils such that the number of tube starts are in proportion to the coil diameter. Thus an outer coil having twice the diameter of an inner coil would have twice the number of tube starts as the inner coil. Note that the effect of this is to keep the length of every tube approximately the same, regardless of which coil it is in. Note there will still be substantial tube length variation when transitioning from a one-start coil to a two-start coil, and similarly from two to three. That is where it is particularly helpful to also include the wire or twisted ribbon insert option in the shorter tubes.

Compactness of the heat exchanger is achieved by two measures: close tube spacing (small gaps, less than one tube diameter, staggered alignment) plus high transfer coefficients that are uniform across the coils (flow through tubes of each coil and around those tubes proportional to the tube surface area of that coil).

Another aspect of achieving compactness involves making the core blocker diameter small compared to the shell diameter. It is generally preferred to size it at less than about half the shell diameter, especially for higher shell-side pressures. The compactness benefit diminishes rapidly below about 20% of the shell diameter.

In order to maintain uniform shell side flow across all sides of all the coils, special provisions are necessary for the innermost and outermost coils. Since they have no adjacent coil on one side, they should have reduced flow in those two locations. That is accomplished by providing smaller gaps at those two locations compared to the coil interior—typically only half as large a gap. Those gaps may be established and maintained by attaching flat spacer strips to the core blocker, and by using tube spacers with reduced coil gaps on the outermost coil.

The greatest overall compactness is achieved when the two gaps are approximately equal. Gaps of about one-half D (0.5 D) have been found to work well for the RHX. As one example of the disclosed heat exchanger applied as an RHX, an apparatus has been built in a 16 inch diameter shell with ten coils of 5/16 inch diameter tubing. The total individual tube count is 34, and the tube lengths vary from 54 feet to 75 feet. The core diameter is 6.3 inches. Tube bundle height is 46 inches.

When two adjacent coils have the same number of tube starts, the respective coil spacers can be interconnected such that the tubes are everywhere in staggered alignment.

However when two adjacent coils have a different number of tube starts, that is no longer possible. There will be one “tube crossing” for each unit of difference in the count, i.e. one place in the circumference where the tubes in the adjacent coils are in-line. Thus those two coils are only partially staggered in alignment, as they are also partially in-line. Accordingly the term here used for a bundle wherein all the adjacent coils with same start count are in full staggered alignment, and all the others are necessarily only partially staggered, is “overall staggered” alignment.

FIG. 7 illustrates how and where the disclosed improved heat exchanger can be applied to an exhaust heat powered ammonia-water absorption cycle which provides chilling to the inlet of a gas turbine, and uses idle capacity to make power. The power is obtained by incorporating ammonia expansion turbine 701 and electric generator 702 into the absorption chilling cycle. The turbine 701 is supplied hot high pressure ammonia vapor via control valve 703, and turbine output is controlled by bypass valve 704. The low pressure expanded ammonia vapor is then absorbed in absorber 322. The disclosed improved heat exchangers can be advantageously applied at any of the refrigerant heat exchanger (RHX) 320; the heat recovery vapor generator (HRVG) 316, and the evaporator/chilling coil 34.

FIG. 8 illustrates another possible application of this invention. It shows an apparatus for extracting useful heat from a hot gas, then cooling it, and then chilling it. It uses at least two and preferably three of the disclosed novel heat exchangers in series in a common containment. Pressure vessel 81 has inlet 82 for the hot gas, and then a first concentric co-coiled heat exchanger 83 for recovering useful heat. Blocker 84 prevents the gas from bypassing the tube bundle 83. A liquid, e.g. ammonia-water solution, is supplied to connector 85 and tube header 86 for distribution to the tubes. That liquid flows upward through the tubes, countercurrent to the hot gas. The heated fluid, e.g. partially desorbed ammonia solution, is gathered into tube header 87, and exits via connector 88. Next the still warm gas is further cooled in a second heat exchanger 89. That exchanger is supplied cooling fluid, e.g. cooling water, via connector 90, and the fluid exits via connector 91. Finally the cooled gas is chilled in the third exchanger, concentric co-coiled bundle 92. A liquid refrigerant, e.g. ammonia, is supplied to that bundle via connector 93, and evaporated refrigerant is removed from that bundle via connector 94.

In some cases the hot gas being cooled will partially condense as it is cooled and/or chilled. That would happen for example when this apparatus is used as the intercooler for an air compression apparatus, or the cooler for a natural gas compressor, or for a compressor of mixed vapors, or an exhaust gas cooler. Disentrainment device 95 directs the larger droplets to a sump, and then the gas goes through demister 96 to the exit port 97. Condensed liquid is removed from the sump while maintaining a liquid seal using for example a float valve 98 actuated by float 99. In an especially preferred embodiment, bundle 83 is the generator (desorber) of an ammonia-water absorption refrigeration cycle, and bundle 92 is the evaporator of that cycle. FIG. 9 is a photograph of a bundle of concentric helical tube coils, showing multiple tube starts in the outer coils, the tube sheet, and the connector welded to the ends of the spacers.