|Publication number||US3921711 A|
|Publication date||Nov 25, 1975|
|Filing date||Mar 20, 1974|
|Priority date||May 30, 1972|
|Publication number||US 3921711 A, US 3921711A, US-A-3921711, US3921711 A, US3921711A|
|Inventors||Adrian J Westbrock|
|Original Assignee||American Standard Inc|
|Export Citation||BiBTeX, EndNote, RefMan|
|Patent Citations (3), Referenced by (26), Classifications (14), Legal Events (2)|
|External Links: USPTO, USPTO Assignment, Espacenet|
United States Patent [191 Westbrock 1 Nov. 25, 1975 1 1 TURBULATOR  Inventor: Adrian J. Westbrock, Birmingham,
 Assignee: American Standard, Inc., New
 Filed: Mar. 20, 1974  Appl. No.: 453,112
Related U.S. Application Data  Continuation of Ser. No. 257,501, May 30, 1972,
 U.S. C1 165/109; 122/367 PF; 122/501; 138/38; 165/174; 165/177  Int. Cl? F28F 13/08; F28F 13/12  Field of Search 138/38; 165/9.1, 9.2, 9.3, l65/9.4, 109 T, 174, 177; 110/97 D; 122/367  References Cited UNITED STATES PATENTS 8/1878 Baker 138/38 X 1,015,131 l/l9l2 Wilson et a1. 122/367 PF 1,953,342 4/1934 Foell l65/9.4
FOREIGN PATENTS OR APPLICATIONS 1,201,074 12/1959 France 122/367 PF 18,542 8/1914 United Kingdom 110/97 D Primary Examiner-Albert W. Davis, Jr.
Assistant ExaminerSheldon Richter Attorney, Agent, or FirmRobert G. Crooks; James J.
 ABSTRACT Turbulator mechanisms for heat exchange tubes in the form of spherical or egg-shaped ball elements packed into each tube so that the tube fluid is forced to flow around the outer surfaces of the ball elements as the fluid moves from the tube inlet to the tube outlet. Successive ones of the ball elements are preferably staggered or offset so that the fluid has multiple directions of turbulence during passage from one ball element to the next.
The invention is particularly applicable for oil coolers wherein the oil admitted to the tubes has a relatively high supply pressure (e.g., 25 to 150 psi), and a relatively low flow rate (e.g. 0.2 to 1.0 gallons per minute per tube). Turbulators of this invention convert a significant portion of the available pressure into useful turbulence, i.e., turbulence that produces a scrubbing action on the tube walls and a correspondingly high coefficient of heat transfer.
1 Claim, 9 Drawing Figures US. Patent ,1 10v.25,1975 Sheet 11 0'52 3,921,711
A m5 m FIG.
Sheet 2 of 2 .385"Bm l s $TAGGERED(F1G.1) .465" BALLS STAGGERED(F1G.1)
VELOCITY .527 1.1). TUBE. wrrH .652." 1.0. TUBE WITH 1,521.0. TUBE WITH .4e55m-| s ALIGNE'D (FM-1.6)
.521 1.0. TUBE (BARE) .521" 1.1). TUBE WITH SPIRAL TURBULATUR (Fla-.7)
c .527"1.o. TUBE WITH A FIN (F16. 5)
US. Patent Nov. 25, 1975 CURVE A m m 7 54 3 2 moor E0 E\ A INNER TUBE SURFACE .2 13 .45 .7 Lo IO 20 30 F169 A P 1 .5.1. (3 FOOT TUBE) TURBULATOR This is a continuation of application Ser. No. 257,501, filed May 30, 1972, now abandoned.
BACKGROUND OF THE INVENTION An investigation of coolers for oil service discloses qualitatively that heat loads are relatively high, flow rates are relatively low, yet moderate pressure drops are permissible. These conditions lead to low Reynolds Numbers for the flow within the tubes, thus laminar flow and low values of heat transfer coefficient.
Multipassing the oil flow improves the situation, increasing the heat transfer coefficient while utilizing more of the available pressure drop; but in most designs excessive numbers of passes are required, increasing exchanger cost while not effectively using the available pressure drop. An optimum design should maximize heat transfer coefficient, using up if necessary the available pressure drop in a single pass. An effective method of turbulation can accomplish this objective.
Another approach for oil cooler design is to use tubes with an internal fin. The additional internal surface decreases total tube length requirements and exchanger size, thus somewhat reducing size and cost. However the internal fins are themselves costly to make and install in the tubes.
I have discovered that spherical elements prove to be excellent core breaker type turbulators in that they quickly convert the available supply pressure into useful turbulence that materially increases the heat transfer coefficient, thus permitting shorter tubes and smaller size coolers for given cooling requirements.
THE DRAWINGS FIG. 1 is a longitudinal sectional view illustrating a portion of a heat exchanger using the invention;
FIGS. 2, 3 and 4 are transverse sectional views taken on lines 22, 33 and 4-4 in FIG. 1;
FIGS. 5 through 7 are transverse sectional views taken through other turbulators and internal fin structures;
FIGS. 8 and 9 are graphs showing the performance of some of the devices depicted in FIGS. 1 through 7;
FIG. 1 shows a portion of a tube-shell heat exchanger comprising spaced tube sheets 10 and 12 having a number of circular heat transfer tubes 14 extending therebetween for conveying oil or other fluid from space 16 to space 18. Spaces 16 and 18 are in practice defined by conventional bonnets or headers (not shown) that direct the fluid to and from the various tubes 14.
The number of tubes 14 and length thereof will vary in accordance with the total mass flow and heat load to be handled by the heat exchanger. Typically the tube length is on the order of three to ten feet, and the tube count is on the order of 100 to 500. Each tube preferably has a diameter of less than one inch. The most economical tube cross section is circular, but the invention is believed to be applicable with other shapes, such as oval shapes.
An object of this invention is to provide turbulator mechanisms in the tubes which will so increase the heat transfer coefficient on the inside surface of each tube as to permit reductions in tube length and tube count. In the FIG. 1 arrangement the turbulators take the form of spherical ball elements 20 packed into the tubes in staggered or offset" relation to one another.
Exterior space 22 between sheets 10 and 12 may be defined or circumscribed by the heat exchanger shell, not shown, said shell being suitably partitioned or baffled by tube-encircling wall elements to direct water or other coolant back and fourth across the outer surface of tubes 14, thereby cooling the oil or other fluid moving through the tubes. The shell and baffles may be of conventional construction.
The coolant in space 22 can alternately be air moving generally transverse to the tubes. In that event the tubes would be provided with external fins, either plate fins or spiral fins, to provide a satisfactory heat transfer at the tube external surface. Ordinarily an air-cooled cooler would be located in a duct system containing a suitable blower for moving the air across the various tubes 14.
As shown in FIG. 1, the spherical turbulator elements 20 are packed relatively tightly into tube 14 so that adjacent ones of the spherical elements are offset or staggered relative to the tube axis and to each other. Each tube 14 is completely filled with spherical elements, from one end to the other, so that the elements maintain their staggered positions whatever the gravitational orientation of the heat exchanger. Spherical elements 20 preferably have diameters somewhat greater than the tube radius in order to achieve the desired staggering and displacement of the fluid away from the core zone toward the tube wall.
Elements 20 are core breaker elements in that they occupy the center or core areas of the tube so as to break up the core zone; i.e., they keep the oil or other fluid in the outer zonesof the tube where the fluid particles can best wash the tube surface.
Core breaker elements 20 can be retained within each tube 14 by any suitable form of retainer. As illustratively shown in the drawing, the retainers consist of coarse mesh screens or grids 24 suitably adhered to the outer surface of the respective tube sheets. Elements 20 can be inserted into the various tubes 14 after one of the screens 24 is adhered to its tube sheet. For example, after screen 24 has been adhered to sheet 10 the tube-sheet assembly can be turned on end, and ball elements 20 can then be poured into tubes 14 until the tubes are completely filled. The other screen 24 may then be adhered onto the tube sheet 12 to retain the balls in place.
During the operation of pouring elements 20 into the tubes the ball elements .will usually automatically assume the illustrated staggered positions because each ball element has a gravitational biasing action on the next lower-most ball element tending to cause the ball elements to assume the most tightly packed condition, i.e., a staggered condition. Should the ball elements for one reason or another not assume the most tightly packed staggered condition it is possible to shake or vibrate the tube-sheet assembly to promote the desired tight packing.
Instead of using loose ball elements it should be possible to employ connected ball elements, i.e., ball elements integrally joined to one another by thin neck sections. For example, the ball elements could probably be molded as a single continuous strip or string having adjacent ones of the ball elements integrally connected together at predetermined points on their peripheries. Assuming a ball diameter of 0.385 inch, the joining neck sections could typically be one-sixteenth inch long and one-sixteenth inch in diameter, sufficient to accomplish the joining function without affecting the turbulating action of the ball elements. The molded strip would of course be cut to the same length as the tube in which it was used. Tubes 14 are copper, admiralty metal, aluminum, or other heat conductive material. Ball elements can be metal, plastic, ceramic, rubber, etc. The balls can be solid or hollow.
FIGS. 2 through 4 illustrate various oil flow areas formed between the inner surface of tube 14 and the outer surface of core breaker elements 20. FIGS. 2 and 4 are taken as transverse sections through the major diameter of two successive elements, while FIG. 3 is taken as a transverse section at a chordal point midway between FIGS. 2-and 4, i.e., the point where adjacent elements contact one another.
While the oil is in the FIG. 2 plane it generally occupies the eastern half of the tube. As the oil moves into the FIG. 3 plane it generally occupies the northern and southern peripheral areas of the tube. Upon reaching the FIG. 4 plane the oil generally occupies the western half of the tube. Similar changes in oil location occur as the oil or other fluid moves downstream from the FIG. 4 plane. This so-called geographical or quadrant change in location of the oil is believed to produce a desired turbulence or shearing action that contributes to a beneficial scrubbing of the oil on the tube surface, such scrubbing in turn being reflected as a higher heat transfer coefficient for the film on the inner side of the tube.
It will be noted that the flow area in a transverse plane passing through the diameter of any one of the ball elements (FIGS. 2 and 4) is somewhat less than the flow area in a transverse plane midway between the ball elements (FIG. 3). Accordingly the linear velocity of the fluid is relatively'fast while the fluid is passing across the ball element diameter and relatively slow while the fluid is at a point midway between adjacent ball elements. The periodic linear velocity changes are believed to promote turbulent mixing of the fluid particles and an advantageous increased scrubbing of the tube surface. With a tube inner diameter of about 0.527 inch and a ball diameter of about .385. inch the linear velocity while in the FIG. 3 plane is calculated to be only about 80 percent of the velocity while in the FIG. 2 or FIG. 4 planes.
The structure depicted in FIGS. 1 through 4 is somewhat idealized in that the ball elements 20 are assumed to be staggered from one another in a regular pattern such that the successive balls are displaced from one another by one hundred twenty radial degrees measured around the tube axis. In actual practice there is no assurance that such a regular displacement will always be achieved. However, because of the fact that the ball diameter is greater than the tube'radius there is an assurance that each ball will roll on the surface of the preceeding ball into a staggered position represent- 3 ing the most compact fill-up of the tube. Thus, successive balls will be staggered in more or less the fashion generally depicted in FIGS. 1 through 4.
It will be noted that the balls have point contact with the tube surface and with each other. The balls thus do not subtract appreciably from the inside tube surface area exposed to flowing oil. This factor is believed to be partly responsible for high heat transfer coefficients achieved.
Another factor contributing to improved heat transfer is believed to be the spherical surface character of elements 20. Except for one point on each spherical surface, the entire surface of each sphere is angled with respect to the direction of fluid flow. Thus, as the fluid flows past a spherical element only the frontal nose of the element is at right angles to the flow: all other surfaces on the sphere are at acute angles to the flow so that such surfaces theoretically produce directional changes in the flowing stream. In theory the frontal nose of a sphere is a point so that the flow directional change area of a sphere is essentially its entire surface area. Since the spheres enjoy point contact with the tube wall and with each other, they present a relatively large total surface to the fluid, hence a high baffling effect.
FIG. 5 shows a cross section through a known heat exchanger tube having a triangular fin comprising three internal chord-like fin walls 30, 32 and 34. The three apex areas of the triangular fin are soldered or otherwise connected to the inner surface of the tube wall so that heat in the flowing oil stream is enabled to travel outwardly through or along the fin material to the apex areas and thence into the tube wall to the surrounding fluid. The triangular fin of FIG. 5 extends the full length of the tube.
FIG. 6 illustrates a heat exchange tube having spherical turbulator elements 20 aligned with one another for the full length of the tube. The general arrangement is similar to the arrangement of FIGS. 1 through 4 except that the spherical elements are not offset or stagger ed from one another.
FIG. 7 illustrates a cross section througha known heat exchanger tube having a spiral fin turbulator 36 graph includes performance curves for the various structures shown in FIGS. 1 through 7. 1
FIG. 9 graphically shows the relation betweenpressure drop and the heat transferred per tube unit surface area. The FIG. 9 curve designations correspond to. the
designations used in FIG. 8. Thus, curve A in FIG. 9 is for the same structure as curve A in FIG. 8. Each curve in FIG. 9 was plotted from four experimental flow values, namely 0.4 gal/min, 0.8 gal/min, 1.2 gal/min and 1.8 gal/min; the experimental values appear as points on the curves.
The following chart illustrates comparative performance for various ones of the tube-turbulator combinations graphed in FIGS. 8 and 9. The chart was prepared using a flow value of 1.8 gal/min (the rightmost point on each curve in FIG. 9).
In the preceding chart the values for Columns 1 and 2 are taken from FIG. 9. Column 1 values are FIG. 9 values divided by three (because the A P in FIG. 9 was for a 3 foot long tube). Thus, for curve A, Column 1 is 0.36 divided by 3, which works out to be 0.12.
Column 3 figures represent the tube footage required to produce a pressure drop of 5 p.s.i. The 5 p.s.i. value is arbitrarily chosen as representative of supply pressures and pressure drops that are commonly available or used in oil cooler installations. The values for Column 3 are calculated by dividing 5 by the Column 1 values.
Column 4 values represent the heat transferred by the tube footage appearing in column 3. Column 4 values are calculated by multiplying the Column 2 and Column 3 values; e.g. 6 X 42 252.
Column 5 figures represent the tube footage required to achieve a given heat transfer (h or q), arbitrarily chosen as 100. Column 5 values are calculated by multiplying the Column 3 values by 100, and then dividing the product by the column 4 values:
mg. 72 X I Column shows the tube footage required to transfer a given quantity of heat from a given mass flow of oil undergoing a given pressure drop (taken as 5 p.s.i.).
The Column 5 figures thus represent the size heat exchanger required when using various different tube-turbulator combinations. It will be seen that the bare tube construction (not turbulator) requires the most tube footage (16 feet), while the staggered ball turbulator construction (curve D) requires the least tube footage (2.2 feet). The curve D construction is therefore the preferred structure.
Curves E and F are somewhat similar constructions, since both employ 0.465 diameter balls in 0.652 diameter tubes. In the curve E structure the balls are staggered or offset, as in FIG. 1, whereas in the curve F structure the balls are aligned with one another, as in FIG. 6. The staggering of the balls appear to improve the performance, since it results in decreasing the tube footage from 6.7 to 2.6. Therefore it is preferred to employ the balls in staggered form as shown in FIG. 1.
It is not known exactly why the staggered ball arrangements give better results than the aligned ball arrangement. It will be seen from FIG. 9 (comparing curves E and F) that for a given flow rate the staggered ball arrangement produces a somewhat higher pressure drop than the aligned ball arrangement; thus if only the pressure drop is considered then the aligned ball arrangement is preferred over the staggered ball arrangement. However, it will be seen from FIG. 8 that the heat transfer coefficient achieved with the staggered ball artube type L L tube footage h for ft. of tube (Curve) ft. tube ft. tube to produce A P to achieve (FIG-9) (FIG. 9) LP of 5 .of5 h of psi, p.s.i.
A .12 6 42. 252 16 B ,3 i 9 l7. 153 1 1 C 1.2 16 4.2 67 6.2 D 13 46 .38 17.5 2.2 E 5 39 1. 39. 2.6 F 1.6 15 3.1 46.5 6.7
rangement is much higher than that achieved with the aligned ball arrangement; on an overall basis the staggered ball arrangement proves to be better than the aligned ball arrangement. The staggering of the ball elements produces a corkscrew directional component that is not present with the aligned ball arrangement. With such a corkscrew component the fluid has a longer residence time in the tube even though its average linear velocity may be somewhat higher (because the staggered balls occupy more of the tube volume than the aligned balls). It is theorized that the longer residence time, coupled with the higher average velocity, may at least partially account for the better heat transfer achieved with the staggered ball arrangement.
Thus far the invention has been tested only on oil coolers wherein oil supply pressures are relatively high (from 15 to 50 p.s.i.), and flow rates are relatively low (e.g., 0.2 to 1.0 gal/min per tube). However, it is known that turbulation can improve the coefficient of heat transfer for air and other gases in boiler tubes, along fin surfaces, and elsewhere. Therefore it is believed that ball turbulators should be useful with fluids other than oil, provided there is sufficient pressure for the turbulators to utilize.
Some thermal situations rule out the use of multipass heat exchangers. Thus, where there is a large temperature differential between the inlet and outlet streams, it is often not feasible to use multi-pass designs because the tubes in the different passes expand or contract different amounts, thus exerting distortional stresses on the tube sheets. By using ball turbulators it is possible to employ single pass heat exchanger designs for these large temperature differential situations.
In the work to date the ball elements in a given tube have been the same diameter. It is believed that useful results could probably be realized with varying diameter balls in the same tube, providing each ball has a diameter greater than the tube radius. The tube I.D. should probably be from about one-half inch to 1 inch; the ball diameter should be in the range from about 0.3 inch to about 0.85 inch. The ball-tube diameter ratio should probably not exceed 0.85.
In tests run on tubes of 0.642 inch diameter and balls of 0.605 inch diameter the pressure drop proved excessive in relation to the improvement in heat transfer coefficient. The ball-tube diameter ratio for this excessive pressure drop arrangement was about 0.94. It is believed that ball-tube diameter ratio between about 0.6 and 0.8 are most suitable or practical. Apparently the balls should have a clearance of about 0.14 to 0.18 inches with respect to the tube wall, measured across the ball diameter. If the clearance, or minimum passage dimension, is decreased below this range there is an excessive pressure drop. If the clearance is increased above this range there is apparently an insufficient turbulence and scouring action on the tube surface, thus an undesirably low heat transfer coefficient.
I claim: 1. A heat exchanger for rapidly conducting heat between a first fluid and a second fluid comprising:
a thermally conductive tubular cylindrical member having a fluid entrant opening at one end and a fluid discharge opening at the other end thereof;
means for a continuous and uninterrupted flow of the first fluid through said tube",
turbulator means disposed within the tubular member for continuous turbulent mixing of the first fluid when the first fluid is being conducted through the heat exchanger from the entrant opening to the discharge opening of the tubular member;
said turbulator means comprising a plurality of spherically formed members being randomly disposed within the tubular member and in which each spherical member is in point to point contact between two adjacent spherical members and in point contact with the inner surface of the tubularmember to therebyprovide three point contact;
said spherical members having a diameter greater than the internal radius of the tubular member and less than the diameter of the tubular member;
said spherical members being positioned along a three-dimensional curve formed by the centroid of each member, the axis of which coincides with the axis of the tubular member, said spherical members being offset from each other by approximately radial degrees measured about the axis of the tubular member and the three-dimensional curve; and
said spherical members forming a path which causes the first fluid to flow around and between the spherical members so that said first fluid substantially contacts the entire inner surface of the tubular member to provide a scrubbing action on the tube walls and when a second fluid is caused to flow around the outer surface of the tubular mem: ber heat is rapidly dissipated through the tube wall so that a heat exchange is accomplished between said first and second fluid.
l I l l
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|U.S. Classification||165/109.1, 138/38, 122/367.4, 165/174, 122/501, 165/177, 165/DIG.530|
|International Classification||F28F13/08, F28F13/12|
|Cooperative Classification||F28F13/12, Y10S165/53, F28F13/08|
|European Classification||F28F13/08, F28F13/12|
|Jun 16, 1986||AS02||Assignment of assignor's interest|
Owner name: AMERICAN STANDARD INC.,
Owner name: ITT CORPORATION, 320 PARK AVENUE, NEW YORK, N.Y.
Effective date: 19860606
|Jun 16, 1986||AS||Assignment|
Owner name: ITT CORPORATION, 320 PARK AVENUE, NEW YORK, N.Y.
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST.;ASSIGNOR:AMERICAN STANDARD INC.,;REEL/FRAME:004561/0629
Effective date: 19860606
Owner name: ITT CORPORATION,NEW YORK
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNOR:AMERICAN STANDARD INC.,;REEL/FRAME:4561/629
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNOR:AMERICAN STANDARD INC.,;REEL/FRAME:004561/0629
Owner name: ITT CORPORATION, NEW YORK