Shell and coil heat exchanger

The heat exchanger is made up of a shell having a coaxial tubular outer and inner wall with end plates attached thereto to enclose a tubular shell cavity provided with an inlet and outlet for a first fluid. Within the shell cavity is a spiral coil of tubing through which flows a second fluid. The coil is wound helically about the axis of the shell and sized to fit the inner and outer walls with limited radial clearance. The coils are axially spaced from one another to define a spiral flow path within the shell cavity for the fluids to first flow. The radial and axial clearance establish a spiral flow path and an axial flow path which are relatively sized to cause the first fluid to travel in a spiral motion, thereby enhancing heat transfer between the first and second fluids.

DESCRIPTION 
1. Field of Invention 
This invention relates to heat exchangers and more specifically shell and 
coil heat exchangers for transferring heat between two fluids. 
2. Background of Invention 
Heat exchangers of a shell and coil design have been used for many years in 
a variety of applications where it is desired to transfer energy between 
two fluids. Shell and coil heat exchangers are frequently used in 
refrigeration systems and heat pumps. Shell and coil heat exchangers can 
be fabricated into a compact unit capable of withstanding relatively high 
pressure. 
Shell and coil heat exchangers are typically mounted vertically, i.e., the 
axis about which the coil is wound is perpendicular to the ground. With a 
vertical shell when you have a gas vapor mixture, the gas will tend to 
accumulate at the top with the shell and the liquid will accumulate at the 
bottom. The flow of the fluid in the shell is generally axial flowing from 
one end to the other and circulating about the coils of tubing within the 
shell cavity. 
In order to minimize the volume within the shell a central tubular insert 
may be provided which falls within the helical coil. This is particularly 
useful in refrigeration systems and heat pumps so that the quantity of 
refrigerant may be minimized. Example of a shell and coil heat exchanger 
having an inner shell to minimize shell cavity volume is shown in U.S. 
Pat. No. 2,668,692 and companion U.S. Pat. No. 2,668,420. In spite of the 
inner shell, a significant disadvantage of the shell heat exchangers are 
the large volume of the shell cavity relative to the volume of the liquid 
within the coiled tubing. 
SUMMARY AND OBJECT OF THE INVENTION 
The object of the invention is to achieve maximum heat transfer rate and 
overall efficiency while minimizing the size of the shell and coil. 
Another object of the invention is to minimize the volume of fluid within 
the shell and coil cavities. 
Another object of the invention is to develop a heat exchanger which 
performs satisfactorily in both the vertical and horizontal positions. 
The present invention is directed to a heat exchanger and method of forming 
same. The heat exchanger is made up of a shell which has a coaxial tubular 
outer and inner wall with end plates attached thereto to enclose a tubular 
shell cavity provided with an inlet and outlet for a first fluid. Within 
the shell cavity is a spiral coil tubing wound helically about the axis of 
the shell and sized to fit between the inner and outer shell walls with 
limited radial clearance. The spiral coil is provided with a plurality of 
windings axially spaced from one another to define a spiral flow path 
within the shell cavity for the first fluid. The radial clearance between 
the spiral coil and shell inner and outer walls is sized such that the 
first fluid travels in a spiral motion to enhance the heat transfer 
between the first fluid and the shell cavity and a second fluid flowing 
within the spiral coil. A dual spiral helical coil assembly for use in the 
heat exchanger may be manufactured using a method made up of the following 
steps: Winding a first tube spirally about a mandrel having a large 
diameter region and a small diameter region. Winding a second tube 
spirally in a similar manner, both tubes having an axial spacing between 
windings. The first and second coils are then interwound so that the small 
diameter region of each coil is nested within the large diameter region of 
the opposite coil. The two coils are then depressed axially to deform the 
coils into a small compact unit with reduced axial spacing between the 
windings. 
The principal advantage of the invention is that the heat exchanger has a 
low enough shell volume so that it works very efficiently in a reverse 
flow heat pump having a heating and cooling cycle. Another advantage of 
the invention is that the fluid within the shell flows in a substantially 
spiral path so that true counterflow can be achieved resulting in maximum 
heat transfer.

BRIEF DESCRIPTION OF THE PREFERRED EMBODIMENT 
With reference to the drawings three preferred embodiments of the heat 
exchanger will be described in detail as well as a method of forming a 
helical coil. 
Embodiment I 
A first embodiment of the heat exchanger is shown in FIGS. 1 through 5. The 
heat exchanger 20 is provided with a pair of spiral coils of tubing 22 and 
24 and the shell assembly 26 made up of the outer tubular wall 28, inner 
tubular wall 30 and first and second end plates 32 and 34. Shell assembly 
26 encloses a tubular shell cavity 36 which is symmetrical about the axis 
of the heat exchanger assembly. At the upper end of the shell assembly is 
a first inlet-outlet fitting 38 and at the opposite end of the shell 
assembly is the second inlet-outlet fitting 40. Each fitting 38 and 40 
communicate with the shell cavity 36 and provide means for admitting and 
means for removing a first fluid from the shell cavity. Which fitting acts 
as an inlet and which fitting acts as an outlet will vary depending on the 
application or the mode of operation in the case of the reverse cycle heat 
pump where the direction of flow may vary depending on whether the unit is 
heating or cooling. 
Within the shell cavity lies the first and second spiral coils 22 and 24. 
Second spiral coil 24 is shown in dotted lines in FIG. 2 so the two coils 
may be distinguished. Both coils are similar shape as shown in FIG. 4. 
Each has a large diameter region 42 and a small diameter region 44 both of 
which are helically wound about the center axis of the heat exchanger 
assembly. The small diameter section of each coil is located within the 
large diameter section of the opposite coil so that the two coils are 
nested together to form a compact assembly. 
The small diameter section of the coil fits relatively closely to the inner 
tubular wall 30 of the shell assembly and the outer periphery of the large 
diameter of the coil fits relatively close to the outer tubular wall 28 of 
the shell assembly. Each of the spiral coils has a plurality of windings 
located generally adjacent a corresponding winding in the opposing coil so 
that the combined radial dimension of the two windings substantially 
occupy the space of the cavity between the inner and outer walls. The 
radial clearance between the inner and outer wall and the pair of coiled 
windings is carefully controlled to restrict the flow of a first fluid 
flowing through the shell cavity. The axial spacing of the coil windings 
is also very carefully controlled to define a spiral flow path within the 
shell cavity for the first fluid. The radial clearance and axial spacing 
of the coils and the shell cavity are relatively sized so that the first 
fluid within the shell cavity travels in a spiral motion to enhance the 
heat transfer between the first fluid within the shell cavity and the 
second fluid traveling within the coils. 
Two spiral coils of tubing are used in the first embodiment shown in FIG. 
1-5 in order to maximize the surface to volume ratio as two tubes with a 
given total cross-sectional area having much greater wall area than a 
single tube of equal cross-sectional area. In typical operation the inlets 
and outlets of coils 22 and 24 will be connected together with a 
"Y"-shaped yoke to provide a single input and a single output. Copper has 
been found to be a preferred material for the spiral coils. Ideally, the 
copper tubing will have its periphery knurled and its internal surfaces 
rifled so that the surface area can be increased. A tube having augmented 
wall surface of this design is described in detail in U.S. Pat. No. 
4,402,359 which is incorporated herein by reference. Tubing with knurled 
exterior and rifled interior is commercially available from Noranda Metal 
Industries, Inc. of Newton, CT. The tube having a knurled exterior is 
particularly advantageous in the present invention in that when a coil 
contacts an adjacent coil or the wall of the shell flow is not completely 
obstructed in the axial direction since fluid can flow between the raised 
knurled protrusions thereby most effectively using the entire heat 
transfer surface of the tubing. Alternatively, S/T TRUEFIN.RTM., an 
augmented finned tube made by Wolverine, P.O. Box 2202, Decatur, Al.35602, 
may be used to form the coils. 
Referring to FIG. 20, there is shown a heat transfer tube 310 having a 
plurality of intergal radially extending pyramid-fins 312 formed in its 
outer surface. The density of the pyramid-fins is between 80 and 500 
pyramid-fins per square inch and the height of the pyramid-fins is between 
0.015 inch for a pyramid-fin density of 500 pyramid-fins per square inch 
and 0.040 inch for a pyramid-fin density of 80 pyramid-fins per square 
inch. The series of threads intersecting each other at 60.degree. so as to 
form a herringbone or diamond pattern. The threads are in the range of 12 
to 30 TPI, preferably about 20 TPI. The heights of the pyramid-fins formed 
is between about 0.037 in at 12 TPI and about 0.015 in at 30 TPI. The 
preferred height of the pyramid-fins is about 0.022 in at 20 TPI. 
When the pyramid-fins are formed on a tube of relatively small thickness, 
the heat transfer enhancement pattern will extend through the thickness of 
the tube wall as shown in FIG. 21 so as to form a doubly augmented tube. 
If the tube wall is thick enough, or if a smooth mandrel is placed inside 
the tube during formation of the external heat transfer enhancement 
pattern, then the inside of the tube will remain smooth. The inside of the 
tube may then be provided with internal fins 314 such as shown in FIG. 22 
of the drawings. These fins may be formed prior to making the outside 
pyramid-fins or at the same time by pressing the tube during knurling onto 
a mandrel placed inside the tube and having suitable grooves for forming 
the fins. The helix angle of the internal fins is between 0.degree. and 
90.degree., preferably between 15.degree. and 45.degree. with respect to 
the longitudinal axis of the tube. 
FIG. 6 shows a sectional view of the spiral flow path formed between the 
tubes and inner and outer shell walls. The axial spacing of the tube coils 
is shown as dimension Y and the spacing between the inner and outer shell 
walls is shown as dimension X. The cross-hatched area defining the spiral 
flow path is an area equal to X times Y minus twice 
the tube area, i.e., X * Y -.pi.DT.sup.2 /2 where DT equals the tube 
diameter. The minimum axial flow area in the shell is equal to the area of 
the shell minus the area of the tubes in the plan view. As shown in FIG. 5 
the axial flow path consists of three small circular paths. The clearance 
between the two tubes and between the tubes and shell wall is shown 
enlarged in FIGS. 3 and 5 for ease in understanding. The actual axial 
clearance between the coils and the wall may be 0.005 inches or less, 
therefore, the axial flow area can be approximated by multiplying the 
axial clearance times the perimeter of each of the circular flow paths so 
that the minimum axial flow area will equal AX*3.pi./2(D1+D2) where D1 
equals the outer diameter of the inner tubular wall, D2 equals the inner 
diameter of the outer tube wall and AX equals the axial clearance. The 
actual axial clearance may be slightly greater than that described by the 
preceding equation since the outer periphery of the coil is knurled or 
finned thereby giving it a slightly smaller effective diameter than that 
measured across the outside diameter of the tube. In Example 4 below, the 
calculated axial clearance is zero since the tubes fit line to line within 
the shell. Even in that extremely tight example there will be some axial 
flow between the knurls or fins thereby allowing effective utilization of 
the entire tube surface area. 
In order to achieve a significant spiral flow path for the first fluid in 
the shell cavity, the axial flow area should not exceed that of the spiral 
flow path as previously calculated. The relationship between the actual 
flow area and the spiral flow path can be quantified by an axial clearance 
ratio which is equal to the axial flow path divided by the spiral flow 
path area. It is therefore desirable to have an axial clearance ratio 
below one hundred percent. It is preferred that the axial clearance ratio 
be maintained below sixty percent. The most preferred axial clearance 
ratio be between zero to sixty percent depending upon the specific 
application for the heat exchanger unit. Note that even with the zero 
axial clearance ratio as previously calculated, there will be some axial 
flow due to the knurling of the coil tubing. The following examples 
represent possible heat exchanger embodiments, the first of which has been 
tested and performed quite satisfactorily. 
EXAMPLE 1 
______________________________________ 
Coil Design Type I 
______________________________________ 
X 1.515 
Y .9375 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.537 
D.sub.1 6.000 
D.sub.2 2.970 
Axial Clearance (AX) 
.005 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.211 
Axial Clearance Ratio 
39% 
______________________________________ 
EXAMPLE 2 
______________________________________ 
Coil Design Type I 
______________________________________ 
X 1.5195 
Y .9375 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.542 
D.sub.1 6.000 
D.sub.2 2.961 
Axial Clearance (AX) 
.0075 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.317 
Axial Clearance Ratio 
58% 
______________________________________ 
EXAMPLE 3 
______________________________________ 
Coil Design Type I 
______________________________________ 
X 1.512 
Y .9375 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.535 
D.sub.1 6.000 
D.sub.2 2.978 
Axial Clearance (AX) 
.004 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.169 
Axial Clearance Ratio 
32% 
______________________________________ 
EXAMPLE 4 
______________________________________ 
Coil Design Type I 
______________________________________ 
X 1.50 
Y .9375 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.523 
D.sub.1 6.00 
D.sub.2 3.00 
Axial Clearance (AX) 
0 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
0 
Axial Clearance Ratio 
0 
______________________________________ 
USE OF HEAT EXCHANGER IN DURAL MODE HEAT PUMP 
The heat exchanger described of the first embodiment works quite 
satisfactorily in a water source heat pump which can be used for both 
heating and cooling. A schematic diagram of a heat pump in the heating 
mode and the cooling mode are shown in FIGS. 7 and 8 respectively. The 
heat exchanger is depicted by box 20 and is provided with water inlet 60 
and water outlet 62. The water circulates through the tubular coil in the 
heat exchanger unit. In the shell of the heat exchanger is circulated a 
refrigerant such as Freon.RTM. 22. In the heating mode, refrigerant enters 
in the outlet 64 and exits the shell cavity through inlet/outlet 66 as the 
refrigerant circulates in the direction of the arrows. The refrigerant is 
circulated by pump 68 which circulates the Freon.RTM. in a closed loop 
path through tube and shell heat exchanger 20, tube and fin heat exchanger 
70. Heat exchanger 70 transmits energy between the Freon.RTM. and air 
which is circulated through the heat exchanger by a blower which is not 
shown in the heating mode and reversing valve 72 and is oriented out that 
the output of the pump is connected to the tubing vent heat exchanger 70 
and the suction side of the pump is connected to a shell and coil heat 
exchanger 
In the heating mode the shell and coil heat exchanger acts as an evaporater 
and the tube and fin heat exchanger 70 acts as a condenser. The hot high 
pressure output of pump 68 flows to tube and fin heat exchanger 70 and is 
cooled by the flow of air therethrough. Pressure is maintained relatively 
high and the tube and fin 70 by expansion valve 74. When the refrigerant 
flows through expansion valve 74, pressure drops substantially. As a low 
pressure refrigerant flows into the heat exchanger 20, it absorbs heat 
from the water circulating through the coils and evaporates. Refrigerant 
exits the heat exchanger through outlet 66 and passes through reversing 
valve 72 to the inlet of pump 60 to complete the heating cycle. 
Pump 60 is driven by conventional mechanical means such as an electrical 
motor. Since heat energy is being added or removed from the water 
circulating through the coil of the heat exchanger, the energy output to 
the air substantially exceeds the energy consumed by the pump 68 in the 
heating and cooling modes. In the cooling mode, the reversing valve 
switches as shown in FIG. 8 so the suction side of the pump is connected 
to the tube and fin heat exchanger 70 and the outlet of the pump is 
connected to the shell and coil heat exchanger 20. In the cooling mode the 
heat exchanger 20 acts as a condenser. The water circulating through the 
coil cools the refrigerant circulating through the shell cavity. The 
refrigerant flows through expansion valve 74 and evaporate in the tube and 
fin heat exchanger 74 to cool the air flowing therethrough. 
It has been determined that the heat exchanger of the present design 
performs quite well in a reverse cycle water source heat pump and is 
capable of achieving very high efficiency levels in both the heating and 
cooling modes. Previous heat pump designs tended to optimize performance 
in one mode that was used most frequently and accepting a lower 
coefficient of performance in the lesser used mode. 
EMBODIMENT II 
An alternative embodiment of the heat exchanger is shown in FIGS. 9 through 
13. In the second embodiment the heat exchanger assembly 80 is provided 
with a first and second spiral coil 82 and 84 helically wound about a 
central axis and having a constant uniform diameter. The two coils are 
interwoven like a double lead screw as shown in FIG. 11. Each of the 
individual coils has substantial axial spacing between the plurality of 
windings as shown in FIG. 9. The coils are identical in structure. The 
shell assembly 86 is made up of an outer tubular wall 88 and an inner 
tubular shell wall 90 which are connected by first and second end plates 
92 and 94 to define a shell cavity 96. A shell cavity is provided with a 
first and second inlet/outlet fitting 98 and 100 at opposite ends of the 
shell cavity. 
A fragmentary cross-sectional side view of a portion of the heat exchanger 
assembly is shown in FIG. 12. The inner and outer walls of the shell 90 
and 88 are spaced apart by a distance slightly greater than the diameter 
of the coils 82 and 84 thereby providing axial clearance for the flow of 
the first fluid in the heat exchanger shell. In FIG. 11 coil 84 is drawn 
in dotted lines to more clearly show that each coil winding is positioned 
between the windings of the other coil. The spiral flow path in the second 
embodiment of the invention is shown in FIG. 13. Note dimension wide the 
distance between coil windings represents the distance between two 
windings of the same coil. The equation defining the spiral flow area is 
the same for the second embodiment as it is for the first. The spiral flow 
area equals X times Y minus D.sup.2 /2. The minimum axial flow area is 
equal to the axial clearance between the tube and shell wall times the 
total clearance area length, i.e., axial clearance area equals pi times 
axial clearance times (D 1+D 2). The following are examples of potential 
designs for heat exchangers of the type shown in the second preferred 
embodiment: 
EXAMPLE 5 
______________________________________ 
Coil Design Type II 
______________________________________ 
X .760 
Y 1.6875 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.400 in.sup.2 
D.sub.1 6.000 
D.sub.2 4.480 
Axial Clearance (AX) 
.005 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.165 
Axial Clearance Ratio 
41% 
______________________________________ 
EXAMPLE 6 
______________________________________ 
Coil Design Type II 
______________________________________ 
X .763 
Y 1.6875 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.405 
D.sub.1 6.000 
D.sub.2 4.477 
Axial Clearance (AX) 
.0065 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.208 
Axial Clearance Ratio 
51% 
______________________________________ 
EXAMPLE 7 
______________________________________ 
Coil Design Type II 
______________________________________ 
X .758 
Y 1.6875 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.396 
D.sub.1 6.000 
D.sub.2 4.484 
Axial Clearance (AX) 
.004 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
.128 
Axial Clearance Ratio 
32% 
______________________________________ 
EXAMPLE 8 
______________________________________ 
Coil Design Type II 
______________________________________ 
X .760 
Y 1.6875 
DT .750 
Spiral Flow Path Area 
X*Y - .pi.DT.sup.2 /2 = 
.400 
D.sub.1 6.00 
D.sub.2 4.50 
Axial Clearance (AX) 
0 
Axial Clearance Area 
.pi.AX(D.sub.1 + D.sub.2)3/2 
0 
Axial Clearance Ratio 
0 
______________________________________ 
METHOD OF WINDING A COIL AND FORMING HEAT EXCHANGER 
FIG. 14 shows a diagram of a mechanism specifically designed for winding 
heat exchanger coils. The apparatus has a central mandrel 110 having a 
large diameter section 112 and a small diameter section 114. The mandrel 
is provided with a helical semi-circular groove having the same large and 
small diameter and the same number of turns to get the coil employed in 
the first embodiment of the invention as shown in FIG. 4. The axial 
spacing between the grooves where the pitch of the spiral on the mandrel 
is significantly greater than the finished coil shown in FIG. 4. The 
semi-circular group 116 corresponds in diameter in the tube size to be 
formed into a coil. 
Mandrel 110 is pivotably supported one end by bearings 118 and 120. The 
mandrel is driven by hydraulic motor 122 which is coupled to the mandrel 
by sprockets 124 and 126 and chain 128. Bearings 118 and 120 and the 
hydraulic motor 122 are affixed to an assembly 130. Affixed to frame 130 
are guide rods 132(a) and 132(b) preferably four parallel guide rods are 
parallel to the axis of the mandrel 110. Sliding axially along the guide 
rods is subframe 134 which is shown in its left most position in FIG. 14. 
Mounted on subframe 134 is guide roll 136 and 138 which are pivotably 
mounted on the ends of hydraulic cylinders 140 and 142. The small end of 
the mandrel 110 is privotably supported by the bearing 144 which is 
affixed to the end of the link 146. Link 146 is pivotably affixed to frame 
130 so that it can be hinged into and out of cooperation with the mandrel 
110 as shown by the arrow in FIG. 14. 
Prior to the bending of a coil, a straight length of copper tube of 
sufficient length to form a coil is selected and filled with sand. The 
ends of the tube are capped to prevent the sand from escaping. The sand 
prevents the tube from kinking or collapsing during the bending process. 
With some thick wall tubing sand is not required. Hydraulic cylinders 140 
and 142 are not fully retracted so that guide rollers 136 and 138 are in 
contact with the mandrel. In the embodiment shown the mandrel would be 
rotated 180.degree. so that clamp 148 would be on the top of the mandrel. 
One end of the tube would then be affixed to the mandrel with clamp 148 so 
that the clamp would be lying in a semi-circular helical group 116. 
Hydraulic cylinders 140 and 142 would then be pressurized causing the 
guide rollers to come in contact with the mandrel. Note that guide roller 
136 is provided with a semi-circular groove to cooperate with a tube to be 
bent. The load exerted by hydraulic cylinders 140 and 142 is substantially 
equal so that there is minimal bending force exerted on the mandrel. With 
the tube clamped in place and the guide rolls in position, hydraulic motor 
122 is activated to cause the mandrel to rotate counter-clockwise when 
viewed from the end adjacent the hydraulic motor. As the mandrel rotates 
the entire subframe assembly 134 with the guide rolls and hydraulic 
cylinders mounted thereon moves to the right in FIG. 14 traversing the 
length of the mandrel. As the subframe reached the transition from the 
large mandrel end 112 to the small mandrel end 114 the hydraulic cylinders 
140 and 142 maintain the guide rolls in constant contact with the mandrel. 
When the desired number of winding have been made, the hydraulic motor 
stops, hydraulic cylinders are retracted and link 146 is pivoted clockwise 
out of the way. Clamp 148 is holding the coil in place is removed and the 
hydraulic motor is run with the coil restrained from turning so that the 
formed coil is screwed off of the mandrel. The formed coil as shown in 
FIG. 15 is substantially longer than ultimately desired and the axial 
spacing between the windings is large. The ends of the coil is then 
uncapped and the sand removed. A second coil is then formed in the 
identical manner so that the two coils are placed end to end with the 
small ends of each coil in contact with one another. The one coil is then 
rotated so that the two coils threadingly interweave with one another so 
that the small end of one coil become located entirely within the large 
end of the opposite coil and vise versa. 
With the two coils oriented in nested relationship with one another as 
previously described, they are then pressed to the desired final length 
using a fixture shown in FIG. 16. The inner shell tubular wall is cut to 
length and welded to the lower end plate to form inner tube end plate 
assembly 160. Assembly 160 is placed on a flat surface and guide mandrel 
162 is telescopingly inserted therein. The lower end of guide mandrel has 
a cylindrical section to fit into the inside diameter of assembly 160 and 
the opposite end of guide mandrel is conically tapered. The inner tube end 
plate assembly with the guide mandrel installed has an overall length in 
excess of the length of the coil spring prior to compression. A coil 
spring pair interwoven as previously described is placed over the guide 
mandrel top place 164 is placed thereon and compressed by Ram 166 using a 
conventional press (not shown). When the top plate 164 has been pressed to 
the inner tube end plate assembly, the top plate is then tack welded to 
the inner tube then the ram and the guide mandrel are removed so that the 
weld can be completed resulting in a spool-like assembly. 
The spool-like assembly 168 which consists of an inner tube top and bottom 
plates and the coils are then fitted with the outer shell walls as shown 
in FIG. 17. The outer shell walls are made up of two identical 
semi-cylindrical halves 170 which are provided with a slot 172 through 
which the ends of the coils may project in an inlet/outlet fitting 174. 
The two semi-cylindrical halves are welded to the top and bottom plates 
and to each other. Yokes 176 (a) and 176 (b) are then welded to the tubes 
projecting through slot 172 and through the shell in a leak-tight manner. 
Note that the yokes used have individual outlets for each of the tubes 
forming the coil assembly, however, it may be more convenient in some 
instances to have a single outlet. With the yokes welded on the unit comes 
complete and it is then pressure tested for leaks and attachment brackets 
as desired are affixed to the outer shell. 
The semi-cylindrical shell halves 170 employed in the preferred embodiment 
of the invention are constructed of steel tubing which has been cut and 
split. The tubing has an 1/8 inch nominal wall thickness and it is 
relatively easy to fabricate and weld. In high volume production, it is 
envisioned that the shell halves could be stamped or rolled with the yoke 
integrally formed therein. 
EMBODIMENT III 
Another alternative embodiment of the invention is shown in FIGS. 18 and 
19. This third embodiment 180 consists of a lower shell and coil heat 
exchanger assembly 182 and upper shell and coil heat exchanger 184 and a 
central receiver 186. The lower shell and coil heat exchanger 182 is 
similar in construction to the first embodiment shown in FIGS. 1 through 6 
and previously described. The upper shell and coil heat exchanger 184 is 
mounted coaxially with the lower shell and coil heat exchanger 182 and 
utilizes a common outer tubular wall 188 and a common inner tubular wall 
190. The third embodiment of the invention is provided with a top and 
bottom endplate, 192 and 194 and a divider plate 196 which separates the 
shell cavity into two independent fluid-tight cavities, upper cavity 198 
and lower cavity 200. Within the lower cavity is a pair of spiral coils 
202 and 204 and within the upper cavity is a single spiral coil 206. 
There are a number of applications when multiple heat exchangers are needed 
in a system and the third embodiment of the invention shown in FIGS. 18 
and 19 provides two heat exchangers in a very small compact assembly. 
Depending on the situation divider plate 196 may be left out thereby 
forming a single shell cavity in which both coil assemblies are housed. 
Heat exchangers of the present design are useful when a desuperheater is 
desired. Desuperheaters are also well known in the art and are used in 
situtations when it is desirable from an efficiency standpoint to reduce 
the presser head pressure by providing supplemental cooling of the 
refrigerant. The top coil is also quite useful in residential dual mode 
heat pump systems where hot water will be heated or preheated by the heat 
pump. In the case of a hot water system or other device used with potable 
water, the coil is formed of a double walled tube for the purpose of 
detecting leaks. Whenever you are using potable water in conjunction with 
a refrigerant, it is important to detect leaks so that Freon.RTM. is not 
introduced into water intended for human consumption. Double wall tube of 
the type made by Noranda Metal Industries, Inc. of Newton, CT 06470 and 
referred to as a leak-detection double augmented tube (LDDA Series) works 
quite satisfactorily when combined with an appropriate leak sensor and 
shut-off or warning system. 
The third embodiment of the invention as shown is also provided with an 
internal receiver 186 defined by inner tube wall 190 and top plate 192 and 
bottom plate 194. Note that unlike a first embodiment of the invention, 
the top and bottom plates enclose the ends of the inner tubular wall to 
form a fluid tight cylindrical cavity. The receiver is provided with an 
inlet 208 and an outlet 210 projected through the top plate 192. Outlet 
210 preferrably is in the form of an elongated tube and extending into the 
receiver cavity and terminating near the bottom thereof. Receivers are 
quite frequently used in refrigerant systems and the present embodiment 
provides a compact receiver with minimal extra cost. It is important to 
note that there will in fact be some heat transfer between the fluid 
contained in the receiver and the fluid in the shell cavity to heat 
transfer through the inner shell wall. This heat transfer can be managed 
in some situations and likewise can be a detriment when no heat transfer 
is desired. When no heat transfer is desired, it is possible to install an 
additional receiver tube slightly smaller in outside diameter than the 
inside diameter of the inner shell wall thereby providing an airgap 
insulation separating the receiver cavity from the rest of the device. 
The coil used in the second embodiment of the invention is somewhat easier 
to fabricate since both coils are uniform in diameter. The apparatus shown 
in FIG. 14 used for the winding of the coil used in the first embodiment 
can also be used to wind the coil in the second embodiment. The mandrel 
110 is provided with a series of axially spaced apart drilled and tapped 
holes 150 for the attachment of clamp 148 at various axial positions along 
the mandrel. As shown in FIG. 14, clamp 148 is attached in the extreme 
leftmost position, a position that would be used for forming a constant 
diameter coil of the type shown in FIG. 9. When a dual diameter coil is to 
be formed of the type shown in FIG. 15, clamp 148 would be attached to the 
mandrel 110 and the center portion of the large diameter region so that 
half of the coil windings will be formed on the large diameter region and 
half on the small diameter region. Two coils are formed with the desired 
number of turns and then they are threadingly fitted into each other. It 
may be necessary to press the unit axially to the desired length to 
achieve a specific axial tube space, however, pressing may not be 
necessary if the axial clearance ratio can be adequately established by 
varying the inside or outside shell wall diameter. 
It will also be understood, of course, that while the form of the invention 
herein shown and described constitutes a preferred embodiment of the 
invention, it is not intended to illustrate all possible forms thereof. It 
will be understood that the words used are words of description rather 
than limitation and various changes may be made without departing from the 
spirit and scope of the invention disclosed.