HEAT EXCHANGER TUBE WITH HIGH HEAT TRANSFER AND LOW PRESSURE DROP

TECHNICAL FIELD

This application relates to heat exchanger tubes, methods of manufacture thereof, methods of using heat exchanger tubes, and methods of heat exchange between fluids incorporating heat exchanger tubes.

BACKGROUND OF THE INVENTION

Heat exchanger tubes are used to provide a thermal transfer between fluids (where the term “fluid” may be gas or a composition in a gaseous or partially gaseous or partially vapor state with or without particulates) for a variety of commercial, industrial, and domestic applications such as hydronic, steam, and thermal fluid boilers, for example. Because of the desire for improved energy efficiency, compactness, reliability, and cost reduction, there remains a need for improved heat exchanger tubes, as well as improved methods of manufacture thereof.

Heat exchanger tubes may be used to convey one fluid from an inlet to an outlet where thermal transfer occurs along its length between a fluid inside the heat exchanger tube to a fluid outside the heat exchanger tube. Resistance to the flow of fluid in an inside of the heat exchanger tube causes a pressure drop from an inlet of the heat exchanger tube compared to an outlet of the heat exchanger tube. This pressure drop represents an undesirable loss of flow pressure that must be overcome by a prime mover (equivalently, a pump, fan or blower) at the cost of energy and system efficiency. Such a prime mover represents a substantial cost—both as an initial investment and as an operating cost—in electricity and fuel expenses, periodic maintenance, and downtime and component replacement costs that typically accompany the requirement for larger, heavier and more costly subsystems and parts. In many industries and applications, the historic remedy for, even avoidable, system thermal inefficiency is larger, heavier and more expensive components and concomitant lifetime operating costs.

Likewise, the dynamics of flow near the boundary layer along both the inside (equivalently, “inner”) surface and outside (equivalently, “outer”) surface of the heat exchanger tube affects the magnitude, location and efficiency of the heat transfer between fluids across the heat exchanger tube wall material.

There remains a need for improved heat exchanger tube design, manufacture and methods for use that can achieve designable, targetable and sustainable low pressure drop, while achieving a high rate of thermal heat transfer, to achieve more compact, efficient and improved heat exchanger systems for commercial, industrial, and domestic applications.

This user requirement is particularly true for retrofit applications for replacement heat exchanger tubes. Heat exchanger tubes are susceptible to blocking, clogging and material failure. Thus, the lifetime of a steam, hydronic or thermal fluid (e.g., cooking oil) heat exchanger typically exceeds the useful life of the originally installed heat exchanger tubes. An important opportunity exists to develop replacement heat exchanger tubes for use in steam, hydronic or thermal fluid applications with the goal of extending the overall lifetime of the heat exchanger system using tubes that can be installed as a normal part of the system maintenance, and offers one skilled in the art of heat exchanger engineering and maintenance the choice of heat exchanger tubes with low pressure drop and high heat transfer characteristics to simultaneously extend the heat exchanger lifetime and improve its operating performance.

More specifically, the inventors have surprisingly discovered novel treatments for heat exchanger tubes that represent significant improvements for a broad spectrum of industries and applications. Heat exchanger tubes are an important point of failure in fluid heating devices; tubes typically have lifetimes shorter than the functional utility of a boiler or heat exchanger, which requires that heat exchanger tubes must be replaced at intervals over the life of the device. Considering retrofit applications together with new heat design applications, without intending to limit the scope or application of the disclosure, three particular situations are contemplated and specifically mentioned where the disclosure presents advancements:

The disclosure provides for heat exchanger tubes in retrofit applications where the requirement is to match the heat transfer rate of the heat exchanger's OEM tube but reduce the overall pressure drop across the replaced tubes to increase system efficiency. Such retrofit replacement tubes preserve the original heat capacity and power density of the heat exchanger design, while extending the system lifecycle and maintenance demands reducing the load on the prime mover.

The disclosure also provides for heat exchanger tubes in retrofit applications where the requirement is to increase the heat transfer rate provided by the heat exchanger tube while approximately maintaining—or even reducing—the pressure drop across the tube, thereby improving the system performance and heat capacity of an existing heat exchanger with little or no penalty in the pressure drop across the heat exchanger. In this situation, the maintenance benefit of retubing the heat exchanger tubes is compounded by an improvement in the performance of the boiler or heat exchanger, while concurrently maintaining or improving the demand requirements on the prime mover (e.g., blower, fan).

The present disclosure further presents new opportunities for one skilled in the art of heat exchanger design to incorporate tubes that exhibit high thermal heat transfer while maintaining low pressure drop for new, compact and efficient heat exchanger and boiler products.

The benefits of the present disclosure are particularly germane in applications—new product and retrofit—where a first fluid is a hot gas mixture (for example, hot combustion gas) flowing inside a heat exchanger tube immersed at least partly in a second liquid fluid (for example, but not restricted to water, water and steam, and steam or oil). Treatment to the inner surface of the exchanger tube is typically most effective because the opportunity to improve the heat transfer rate is high—and can be improved by the current disclosure—since the flow inside the tube is physically constrained and even modest physical enhancements that increase heat transfer from the hot laminar flow along the longitudinal center of the tube, through the boundary layer and, by convection, to the inner wall of the tube are significant. The approximate rate of heat transfer between the bulk of the fluid inside the pipe and the pipe external surface can be expressed as where q is the heat transfer rate (W), where h (or h-factor) is the convective heat transfer coefficient (W/(m2·K)), t is the wall thickness (m), and k is the wall thermal conductivity (W/m·K) A is the surface area (m2) across which heat transfer occurs. In practical applications, the h-factor for the liquid outside the heat exchanger tube is an order of magnitude—sometimes, several orders of magnitude—smaller than the h-factor of the hot gas inside the heat exchanger tube. Thus, in such applications the bulk heat transfer constraint or limiting factor is the gas-side convective heat transfer coefficient. Typically, the second fluid outside the tube is both substantially cooler and denser than the hot gas flowing through the heat exchanger tube, which results in a high heat transfer rate per unit tube length from the hot gas inside the tube, down the temperature gradient, to the second fluid (e.g., production fluid) outside the tube.

SUMMARY OF THE INVENTION

Disclosed herein is a heat exchanger tube with a much higher bulk heat transfer rate and similar pressure drop per unit of heat exchanger tube length relative to a heat exchanger tube with a smooth inner surface of the same dimensions, material construction and operating conditions.

Also disclosed is a pattern altering an inside surface of the heat exchanger tube through scribing, embossing or adding elements that promotes heat transfer while maintaining low pressure drop.

The above described and other features are exemplified by the following figures and detailed description.

BRIEF DESCRIPTION OF THE DRAWINGS

Referring to the figures, which are exemplary embodiments, and wherein the like elements are numbered alike.

FIG. 1A shows a perspective view of a section of a heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 1B shows a perspective view of a straight heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 1C shows a perspective view of a curved heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 1D shows a perspective view of a helical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 1E shows a perspective view of a heat exchanger tube with a compound curve shape in accordance with embodiments of the present disclosure.

FIG. 2[A] shows a schematic of a ridge disposed on the inner surface of a circular cylindrical heat exchanger tube showing the coordinate systems used to describe key disclosure features in accordance with embodiments of the present disclosure.

FIG. 3A shows a cross-section of a spiral ridge with constant spiral angle, elevation angle and constant pitch disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 3B shows a cross-section of a spiral ridge with non-constant spiral angle, elevation angle and constant pitch disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 4A shows a perspective view of a circular cylindrical heat exchanger tube with smooth inner surface and the inlet and outlet pressure measurement points for calculation of the resulting flow pressure drop in accordance with embodiments of the present disclosure.

FIG. 4B shows a perspective view of a spiral ridge with constant elevation spiral angle, elevation angle and constant pitch disposed on the inner surface of a circular cylindrical heat exchanger tube with the inlet and outlet pressure measurement points for calculation of the resulting flow pressure drop in accordance with embodiments of the present disclosure.

FIG. 5A shows a cutaway view of a spiral ridge with a circular corrugation cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5B shows an expanded cutaway view of a spiral ridge with a circular corrugation cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5C shows a cutaway view of a spiral ridge with a solid semi-circular cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5D shows an expanded cutaway view of a spiral ridge with a solid semi-circular cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5E shows a cutaway view of a spiral ridge with a solid semi-elliptical cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5F shows an expanded cutaway view of a spiral ridge with a solid semi-elliptical cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5G shows a cutaway view of a spiral ridge with a solid rectangular cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5H shows an expanded cutaway view of a spiral ridge with a solid rectangular cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5I shows a cutaway view of a spiral ridge with a solid trapezoidal cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 5J shows an expanded cutaway view of a spiral ridge with a solid trapezoidal cross-section disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 6A shows cross-section view illustrating the flow in the boundary layer near the inner heat exchanger tube wall near a ridge with height below the boundary layer transition to the free stream flow regime in accordance with embodiments of the present disclosure.

FIG. 6B shows cross-section view illustrating the flow in the boundary layer near the inner heat exchanger tube wall near a ridge with height extending above the boundary layer transition to the free stream flow regime in accordance with embodiments of the present disclosure.

FIG. 7A shows a cutaway view of for an ensemble trajectory propagation of four distinct flow initial conditions near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 7B shows an expanded cutaway view of an ensemble trajectory propagation of four distinct flow initial conditions near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 7C shows an expanded cutaway view of the relative initial conditions for an ensemble trajectory propagation of four distinct flow initial conditions near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 7D shows an expanded cutaway view of the relative initial conditions for an ensemble trajectory propagation of four distinct flow initial conditions near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube showing the propagation of shear swirl past sequential ridge sections in accordance with embodiments of the present disclosure.

FIG. 7E shows an expanded cutaway view of a short section of the trajectory propagation an initial condition near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube showing the propagation of shear swirl past sequential ridge sections in accordance with embodiments of the present disclosure.

FIG. 7F shows a plot of decaying shear swirl magnitude for an ensemble trajectory propagation of four distinct flow initial conditions near a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 8A shows a perspective view of a pattern of gaps in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 8B shows a cross-section of a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube showing the angular separation of gaps in accordance with embodiments of the present disclosure.

FIG. 9A shows a longitudinal cross-section of a pattern of gaps including an angular offset in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 9B shows a cross-section of a rectangular flow gap in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 9C shows a cross-section of a semi-circular flow gap in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 9D shows a cross-section of a trapezoidal flow gap in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 9E shows a cross-section of a rectangular flow gap with fillet edges in a spiral ridge disposed on the inner surface of a circular cylindrical heat exchanger tube in accordance with embodiments of the present disclosure.

FIG. 10A shows a pair of spiral ridges disposed on the inner surface of a circular cylindrical heat exchanger tube with a separation angle at the points of crossing in accordance with embodiments of the present disclosure.

FIG. 10B shows a cross-sectional diagram of a triplet of spiral ridges disposed on the inner surface of a circular cylindrical heat exchanger tube with constant, equal elevation angles in accordance with embodiments of the present disclosure.

DETAILED DESCRIPTION

As further discussed herein, the Applicants have discovered that heat exchanger tubes can suffer degraded performance through high pressure drop from an outlet of the heat exchanger tube compared to an inlet of the heat exchanger tube due to surface irregularities, whether unintentional (for example, due to deposits, pitting or surface degradation) or intentional (for example, surface treatments such as corrugation designed to improve flow mixing). High pressure drops as measured from the heat exchanger tube inlet to an outlet are responsible for increased cost for fluid drivers (e.g., fans, pumps) required to overcome large pressure losses, and other undesirable costs due to system inefficiencies and increased consumption of fuel, electricity and maintenance resources. Moreover, thermal and pressure drop inefficiencies require larger components (e.g., pumps, blowers, fans and additional tubes) to achieve the design goals which increase subsystem component costs and product footprint. Increased thermal efficiency and lower tube pressure drop enable smaller, more compact designs at lower initial and recurring investment costs.

Moreover, the Applicants have discovered that heat exchanger tubes can suffer suboptimal performance due to poor heat transfer from fluid one to fluid two across the material cross section of the heat exchanger tube wall. The efficiency of heat transfer across the wall of a heat exchanger tube is affected by many factors including the tube material properties (for example, but not limited to, thermal conductivity) and the details of the boundary layer of the fluid flow proximal to the inner and outer surfaces of the heat exchanger tube.

Disclosed is an improved heat exchanger tube incorporating a ridge structure on the inner surface of the tube that provides desirable, designable and targetable low pressure drop and high heat transfer for applications that require heat generation which provides improved efficiency, apparatus lifecycle and performance by alleviating or eliminating the disadvantages described above. The disclosed improved heat exchanger tube is suitable for retrofit applications.

Disclosed are methods for manufacture and application to at least three important examples of applications including retrofit (replacement) heat exchanger tubes that match of the original tube heat transfer rate but lower the pressure drop across the tube; retrofit heat exchanger tubes that increase the heat transfer rate compared to the original tube design, but maintains or lowers the pressure drop across the tube for an improvement in overall thermal and system efficiency performance; and, third, new heat exchanger tube applications for compact, efficient heat exchangers and boilers that can exploit the benefit of high heat transfer and low pressure drop heat exchanger tubes without local boiling or hot spots that could cause material failure.

For example, a boiler is a fluid heating system incorporating a heat exchanger that may be used to exchange heat between any suitable fluids—equivalently, a first fluid and the second fluid—wherein the first and second fluids may each independently be a gas or a liquid or combination thereof. Particularly mentioned is a boiler heat exchanger with a hot gaseous first fluid flowing through the inside of a heat exchanger tube, the tube at least partially immersed in a liquid or liquid/vapor second fluid (for example, but not limited to water, water and steam, or oil) second fluid. Because the gaseous first fluid can be significantly laminar flow through the heat exchanger tube, heat exchanger tube treatments applied to the inner surface of the heat exchanger tube are most effective in increasing the thermal heat transfer rate per unit of tube length by increasing the convection of energy from the laminar free stream flow through the boundary layer whose laminar flow is perturbed by the mixing action (shear swirl, as further described below) induced by the inner wall treatment. The term treatment refers to a modification of the surface designed to improve the performance of the heat exchanger tube. Moreover, preventing laminar boundary layer flow is key to high thermal heat transfer rates, since laminar flow in the boundary layer creates an insulative heat barrier to the convective transfer of heat energy down the temperature gradient to the solid tube wall material. Treatments can be applied to the outside of the tube (for example, fins), but with lesser benefit: the flow of the second fluid outside the tube is not constrained or regular, and the high fluid density limits the effectiveness of small changes in the local heat transfer coefficient. As described above, the bulk heat transfer limiting constraint is the result of a small gas-side h-factor compared to a higher liquid-side h-factor.

In the present Application, the first fluid, which is directed through a heat exchanger core through at least one heat exchanger tube may comprise a liquid, gas or combination thereof that may also include suspended particles. Not to be limited by theory, the first fluid may comprise a combustion gas (for example, a gas produced by fuel fired combustor) and may comprise, for example, water, carbon monoxide, nitrogen, oxygen, carbon dioxide, combustion byproducts or combination thereof. The first fluid may be a product of combustion from a hydrocarbon fuel such as natural gas, oil, wood, propane, or diesel, for example, or gas heated by a heating element such as an electrical conduit, metal wire or resistive conduit, for illustrative examples.

An embodiment in which the first fluid comprises predominately gaseous products from combustion of natural gas or propane, and further comprises liquid water, steam, or a combination thereof and the production fluid comprises liquid water, steam, a thermal fluid, or a combination thereof is specifically mentioned.

Also, the second fluid contacts at least a portion of an outer surface of a heat exchanger tube and may comprise a liquid, gas or combination thereof that may also include suspended particles, such as water, steam, oil, a thermal fluid (e.g., a thermal oil), or combination thereof.

An embodiment in which the second fluid comprises predominately water, steam or a combination thereof is specifically mentioned.

The thermal fluid may comprise water, a C2 to C30 glycol such as ethylene glycol, a unsubstituted or substituted C1 to C30 hydrocarbon such as mineral oil or a halogenated C1 to C30 hydrocarbon wherein the halogenated hydrocarbon may optionally be further substituted, a molten salt such as a molten salt comprising potassium nitrate, sodium nitrate, lithium nitrate, or a combination thereof, a silicone, or a combination thereof. Representative halogenated hydrocarbons include 1,1,1,2-tetrafluoroethane, pentafluoroethane, difluoroethane, 1,3,3,3-tetrafluoropropene, and 2,3,3,3-tetrafluoropropene, e.g., chlorofluorocarbons (CFCs) such as a halogenated fluorocarbon (HFC), a halogenated chlorofluorocarbon (HCFC), a perfluorocarbon (PFC), or a combination thereof. The hydrocarbon may be a substituted or unsubstituted aliphatic hydrocarbon, a substituted or unsubstituted alicyclic hydrocarbon, or a combination thereof. Commercially available examples include Therminol® VP-1, (Solutia Inc.), Diphyl® DT (Bayer A. G.), Dowtherm® A (Dow Chemical) and Therm® S300 (Nippon Steel). The thermal fluid can be formulated from an alkaline organic compound, an inorganic compound, or a combination thereof. Also, the thermal fluid may be used in a diluted form, for example with a concentration ranging from 3 weight percent to 10 weight percent, wherein the concentration is determined based on a weight percent of the non-water contents of the thermal transfer fluid in a total content of the second fluid.

The various components of the heat exchanger tube can each independently comprise any suitable material. Use of a metal is specifically mentioned. Representative metals include iron, aluminum, magnesium, titanium, nickel, cobalt, zinc, silver, copper, and an alloy comprising at least one of the foregoing. Representative metals include carbon steel, mild steel, cast iron, wrought iron, a stainless steel such as a 300 series stainless steel or a 400 series stainless steel, e.g., 304, 316, or 439 stainless steel, Monel, Inconel, bronze, and brass.

Specifically mentioned is an embodiment in which the heat exchanger tube components each comprise steel. Use of a steel, such as mild carbon steel or stainless steel is specifically mentioned. While not wanting to be bound by theory, it is understood that use of stainless steel can help to keep the components below their respective fatigue limits, potentially eliminating fatigue failure as a failure mechanism, and promote efficient heat exchange.

The disclosed system can alternately comprise, consist of, or consist essentially of, any appropriate components herein disclosed. The disclosed system can additionally be substantially free of any components or materials used in the prior art that are not necessary to the achievement of the function and/or objectives of the present disclosure.

The systems and methods have been described with reference to the accompanying drawings, in which various embodiments are shown. This disclosure may, however, be embodied in many different forms, and should not be construed as limited to the embodiments set forth herein. Rather, these embodiments are provided so that this disclosure will be thorough and complete, and will fully convey the scope of the disclosure to those skilled in the art. Like reference numerals refer to like elements throughout.

FIG. 1A shows a drawing of a section of a heat exchanger tube used to convey a first fluid from an inlet to an outlet. A heat exchanger tube may have the shape of any closed curve in a cross-section at each point along the length of the tube including a circle, oval, ellipse, square, rectangle, trapezoid or triangle. A circle is specifically mentioned. For simplicity in illustrating embodiments in the disclosure that follows, diagrams will be provided with heat exchanger tube with circular cross-sectional planar shapes. The outer diameter of a heat exchanger tube 54, denoted D_O, defines a surface 10 exposed, at least partly, to the second fluid. Note that D_Oneed not be constant along the length of the heat exchanger tube. The inner diameter 52, denoted D_I, defines a surface 20 exposed, at least partly, to the first fluid. The inner diameter 52, D_I, also does not need to be constant along the length of the heat exchanger tube. Consequently, the heat exchanger tube wall material thickness 56, denoted THK, need not be constant along the length of the heat exchanger tube. At each point along the length of the heat exchanger tube, heat is transferred from a first fluid flowing 40 through the tube at least partially in contact with the inner surface 20, across the material wall thickness 56, THK, of the tube wall 30, to a second fluid flowing 50 outside the tube partially in contact with the outer surface 10.

A heat exchanger tube may have a three-dimensional shape including, but not limited to, embodiments such as a straight circular cylinder 60 (equivalently, a straight pipe or straight tube) as shown in FIG. 1B, a simple curved circular cylinder 70 (equivalently, a curved pipe or curved tube) as shown in FIG. 1C, a helical circular cylinder 80 (equivalently, a helical pipe or helical tube) as shown in FIG. 1D, or a compound curve circular cylinder 90 as shown in FIG. 1E. A straight circular cylinder (equivalently, a straight pipe or straight tube) 60 is specifically mentioned. By simple methods of analytical geometry known to one skilled in the art of heat exchanger tube design, a heat exchanger tube's shape may be characterized by a three-dimensional curve, S(1), for 0≤1≤L where I denotes a point along the length of the heat exchanger tube curve, S, from zero to the total length, L, of the spatial curve running along the centerline 65 of the heat exchanger tube as illustrated by embodiments such as those shown in FIG. 1B and FIG. 1C. By principles of elementary geometry, the spatial curve, S, can be defined in terms of Cartesian coordinates in a standard way and expressed as the coordinate vector S(l)=(x(l), y(l), z(l)) for 0≤l≤L. (Also shown in FIG. 2.) For example, for the straight tube circular cylinder 60 embodiment shown in FIG. 1B,

S(l)=(x(l),y(l),z(l))=(c₁,c₂,c₃·l) for 0≤l≤L

where c1, c2 and c3 are constants. Other equivalent coordinate representations are possible, and the conversions are known to one skilled in the art of heat exchanger tube design.

The overall length, L_O, of a heat exchanger tube may be different that the spatial curve length, L. Note that for a straight circular cylinder heat exchanger tube 60 as illustrated in FIG. 1B, L_Oequals L. For heat exchanger tubes characterized by different spatial centerline curves, S, L_Odoes not necessarily equal L. For example, for the embodiment of a simple curved heat exchanger tube 70 with a small curvature as shown in FIG. 1C, L>L_O. Embodiments of heat exchanger tubes where L>>L_Oyield compact heat exchanger designs such as those utilizing helical heat exchanger tubes 80 as illustrated in the embodiment shown in FIG. 1D. Repeated bends in a compound curve 90 heat exchanger tubes such as illustrated in FIG. 1E may also be utilized in compact heat exchanger designs that fit in constrained spaces such as within flat or narrow boiler pressure vessels.

The inventors have surprisingly discovered that a ridge structure disposed on the inner surface of a heat exchanger tube, where the dimensions of the ridge structure are judiciously selected, can achieve heat exchanger tube embodiments that exhibit both a low pressure drop and, simultaneously, high thermal heat transfer between a first fluid and second fluid.

FIG. 2 illustrates an embodiment of a ridge 200 disposed on the inner surface 210 of a heat exchanger tube 220. A ridge 200 may have any shape that can be disposed on the inner surface of a heat exchanger tube such as a helix, spiral, straight or curved ridge. A spiral is specifically mentioned and both left-handed and right-handed spiral ridges are contemplated. A helix (alternatively, helical) ridge is specifically mentioned and both left-handed and right-handed helical ridges are contemplated although, not to be bound by theory, the ridge need not have the precise shape of a mathematical helix as both approximate and equivalent helical shapes are specifically contemplated. The inner surface of the heat exchanger may have a plurality (one or more) ridges disposed on the inner surface. One spiral ridge is specifically mentioned. If a plurality of spiral or helical ridges are disposed on the inner surface of the heat exchanger, the individual ridges in an N-helical or N-spiral structure may be denoted Hj or Rj for 1≤j≤N where N denotes the number of ridges comprising the heat exchanger inner surface treatment structure. This nomenclature can be extended to other composite ridge treatment structures to identify the individual components.

It will be convenient in the present disclosure to use various coordinate systems to describe shapes, aspects and features of heat exchanger tubes and disposed ridges used to enhance thermal heat transfer while maintaining low pressure drop. FIG. 2[A] illustrates the coordinate systems useful in the present disclosure. The standard Cartesian coordinate system with coordinates (x, y, z) introduced above is useful for describing the three-dimensional shape of the heat exchanger tube, S(l)=(x(l), y(l), z(l)), as discussed above. This is also useful for describing the three-dimensional shape of an individual ridge in Cartesian coordinates, R(l)=(x_R(l), yR(l), zR(l)), notation that can be further indexed for a compound heat exchanger tube treatment comprised of a plurality of ridge structures. Simple mathematical relationships may be used to convert these Cartesian coordinates into cylindrical coordinates, R(l)=(rR(l), ωR(l), zR(l)), known by one skilled in the art of heat exchanger tube design.

It is also convenient to define coordinates (t, r, l), along the path of the local to each point 255, p, along the ridge space curve, R. These coordinates allow for a description of features that are incorporated along the length of the ridge—for example, gaps that occurring regularly along the ridge spatial curve. The (t, r, l) coordinates are defined by the structure of the ridge and change orientation as the point, p, moves along the ridge space curve where the surface of the ridge contacts the inner surface of the heat exchanger tube. In particular, is tangent to the ridge where it contacts the heat exchanger tube. (That is, t(p)=t′(p)/∥t′(p)∥ where t′ is the usual nomenclature for the spatial derivative of the ridge space curve and ∥·∥ is the norm or length of the spatial vector derivative.). Likewise, using standard formulas, r at p is the coordinate vector normal to the contact surface between the inner surface of the heat exchange tube and point p, and A is the right-hand coordinate vector orthogonal to t(p) and r(p).

For example, FIG. 3A shows an embodiment of a circular cylindrical heat exchanger tube 30 with a single spiral ridge 305 disposed on the inner surface 20 of the heat exchanger tube 30 with external surface 10. In the embodiment shown, the heat exchanger tube inner diameter, D_I, is equal to 10.21 millimeters (mm); equivalently, 0.4 inches. The heat exchanger tube outer diameter, D_O, is equal to 12.7 mm; equivalently, 0.5 inches. Thus, the heat exchanger wall thickness, THK, is 1.25 mm; equivalently, 0.049 inches. The overall length of the heat exchanger tube, L_O, is equal to total spatial curve length, L, measured to be 406.4 mm; equivalently, 16.0 inches.

For a heat exchanger tube with inner diameter, D_I, outer diameter D_Oand overall length, L_O, the shape of a spiral ridge 305 can be parameterized by any one of three parameter choices: the pitch 308, P_s, the spiral angle 309, β, or the elevation angle, α. The pitch of a single spiral is the distance between adjacent spiral section along the longitudinal axis which may be constant or, alternatively, may be a function of the ridge's position along its spatial curve. The spiral angle, β, is the angle between the spiral and a cross-section of the heat exchanger tube. The spiral elevation angle, α, is the angle between the spiral and the longitudinal axis. These parameterizations are related by simple formulae:

β=π/2−a (equivalently, β=90 deg−α)

P
_s
=π·D
_I·tan(β)

In the exemplary embodiment illustrated in FIG. 3A, the pitch angle, Ps, is 12.65 mm (equivalently 0.5 inches) and constant; consequently, the spiral angle is 21.54 degrees, and the elevation angle is 68.46 degrees.

It has been surprisingly discovered that the spiral angle (correspondingly, the pitch and the elevation angle), need not be constant along the length of the ridge disposed on the inner surface of the heat exchanger tube and a variable spiral pitch may be used to affect the resulting heat exchanger tube pressure drop and heat transfer characteristics. In fact, a non-constant spiral angle (equivalently, non-constant elevation and non-constant pitch) can be exploited by one skilled in the art of heat exchanger tube design to match the beneficial mixing induced in the boundary layer flow by the ridge to the changing properties of the flow as it traverses the length of the heat exchanger tube. As the fluid flow migrates down the tube, heat energy is extracted from the flow which lowers the flow temperature, increases the fluid density and decreases the fluid velocity—particularly in a compressible first fluid. These fluid property changes affect the generation of swirl behind the ridge, a generation mechanism that is determined by the ridge geometry, including the spiral angle. FIG. 3B illustrates an embodiment of a single spiral ridge 310 disposed on the inner surface 20 of the heat exchanger tube 30 with non-constant pitch (equivalently, non-constant spiral angle and non-constant elevation angle) and outer surface 10. In this embodiment, the pitch, Ps, increases along the length of the heat exchanger tube so that at two distinct points, s₁311 and s₂312, along the heat exchanger tube spatial curve, the corresponding pitch satisfies P_s1≠P_s2. For example, in the embodiment shown in FIG. 3B, P_s1<P_s2. Likewise, the corresponding elevation angles, α₁>α₂and spiral angles, β₁<β₂also change at each point along the length of the ridge disposed on the inner surface 20 of the heat exchanger tube 30 according to the formulae equivalent to those prescribed above.

It has also been surprisingly discovered that disposing one or a plurality of ridges on the inner surface of a heat exchanger tube can improve the heat transfer between a first fluid flowing inside the heat exchanger tube to a second fluid outside the heat exchanger tube, relative to a smooth tube, while maintaining a low differential increase in the pressure drop from the inlet to the outlet of a heat exchanger tube. To quantify the improvement induced by ridges for design purposes, the inventors have discovered that it is useful to compare—at the same inlet flow velocity—key performance parameters for cases where, (a) the inner surface of the heat exchanger tube has been augmented with the disposition of one or a plurality of ridges comprising specific geometrical features; in compared to, (b) the corresponding smooth heat exchanger tube of the same dimensions without inner tube wall treatment. FIG. 4A shows an embodiment of a straight circular cylindrical heat exchanger tube with outside diameter, D_O, inside diameter, D_I, with corresponding wall thickness, THK, useful for defining these measures, although the same definitions apply to heat exchanger tubes with other geometries including those described above for the embodiments shown in FIG. 1A through FIG. 1E.

For the embodiment illustrated in FIG. 4A, we define the pressure drop, ΔPr_AB, across the heat exchanger tube 30 with outer surface 10 to be the magnitude of the difference between the fluid pressure, P_rB, measured at the heat exchanger tube outlet 410 (labeled “B”) minus the fluid pressure, P_rA, measured at the heat exchanger tube inlet 400 (labeled “A”). Elementary fluid flow principles requires P_rA, >P_rBsince flow energy is dissipated by friction as the fluid flows 420 through the heat exchanger tube from the inlet 400 to the outlet 410. This friction is due, in part, to shear along the boundary layer formed where the flow contacts the inner surface 20 of the heat exchanger tube 30. The pressure loss is cumulative as the fluid flow along the length of the heat exchanger tube which, for the cylindrical heat exchanger tube 30 implies L=L_O. The energy dissipation due to fluid flow friction and, hence, the pressure drop ΔPr_AB, across the heat exchanger tube 30 depends upon several factors including the heat exchanger tube material properties, the flow velocity 420 and the inner surface characteristics (e.g., roughness, etc). Where all other material and flow properties are the same, the pressure drop ΔPr_AB, across a heat exchanger tube 30 with a smooth inner surface is expected to achieve a minimum value.

Compare the smooth tube case with the embodiment illustrated in FIG. 4B, where a ridge 310 has been disposed on the inner heat exchanger tube surface 20. As before, the pressure drop, Δ√{square root over (PR_AB)}, across the heat exchanger tube 30 with outer surface 10 to be the magnitude of the difference between the fluid pressure, √{square root over (Pr_B)}, measured at the heat exchanger tube outlet 410 (labeled “B”) minus the fluid pressure, P_rA, measured at the heat exchanger tube inlet 400 (labeled “A”). Note that while the inlet pressure, P_rA, is the same as from the smooth exchanger tube case shown in FIG. 4A, the outlet pressure, √{square root over (Pr_B)}, is different since the pressure drop is expected to be different. As before, elementary fluid flow principles require that P_rA, >√{square root over (Pr_B)} since flow energy is dissipated by friction as the fluid flows 420 through the heat exchanger tube from the inlet 400 to the outlet 410. This friction is due, in part, to shear along the boundary layer formed where the flow contacts the inner surface 20 of the heat exchanger tube 30, friction that is typically greater since any treatment of the inner tube surface—including the addition of a ridge—represents an obstacle that can increase the flow friction and, hence, pressure drop. Thus, where all other material and flow properties are the same, the pressure drop Δ√{square root over (Pr_AB)} across a heat exchanger tube 30 with an inner surface treatment (e.g., ridge) is expected to be higher than in the corresponding smooth tube case.

Using the convention described above, the following measures have proven to be useful: For a first fluid flowing in the heat exchanger tube surrounded by a second fluid, the heat transfer coefficient, h, is the rate of heat transfer between a solid surface and a fluid per unit surface area per unit temperature difference. The heat transfer coefficient depends on the fluid's physical properties and the physical geometry. Then if h_enhancedis the heat transfer coefficient of a heat exchanger tube enhanced by the disposition of one or a plurality of ridges on the tube inner surface, and h_smoothis the heat transfer coefficient of a smooth heat exchanger tube not enhanced by the disposition of one or a plurality of ridges on the tube inner surface, then we define:

HFR=h_enhanced/h_smooth

as the ratio of heat transfers for treated and untreated heat exchanger tubes at the same inlet fluid velocity and ambient pressure and temperature. Similarly, if ΔP_enhancedis the pressure drop from the heat exchanger tube inlet to the heat exchanger tube outlet for a heat exchanger tube enhanced by the disposition of one or a plurality of ridges on the tube inner surface, and if ΔP smooth is the pressure drop from the heat exchanger tube inlet to the heat exchanger tube outlet of a smooth heat exchanger tube not enhanced by the disposition of one or a plurality of ridges on the tube inner surface, then we define:

DPR=ΔP_enhanced/ΔP_smooth

as the ratio of pressure drops for treated and untreated heat exchanger tubes measured for the same inlet fluid velocity and ambient pressure and temperature.

Using these conventions, a key novelty result can be restated:

The inventors have surprisingly discovered that a ridge structure disposed on the inner surface of a heat exchanger tube, where the dimensions of the ridge structure are judiciously selected, can achieve heat exchanger tube embodiments that exhibit both a low pressure drop increase compared to the smooth tube case (equivalently, DPR close to one, DPR≈1) and, simultaneously, a substantial increase in the thermal heat transfer between a first fluid and second fluid, HFR>>1.

The inventors have also surprisingly discovered that a ridge structure disposed on the inner surface of a heat exchanger tube, where the dimensions of the ridge structure are judiciously selected, can be used to achieve heat exchanger tube embodiments that match the heat transfer characteristics of an inferior treatment or smooth tube, but exhibits a lower pressure drop across the heat exchanger tube length compared to an existing tube with an inferior treatment. This has particular utility in retrofit applications where the original heat exchanger tube design energy transfer rate target must be maintained, but a lower pressure drop is desirable.

There are several important aspects to the inventor's discoveries. A first aspect is that the geometry of the ridge structure and the geometry of its disposition on the inner surface of the heat exchanger tube provide important design parameters that determine the resulting HFR and DPR of the treated heat exchanger tube. FIG. 5A shows an embodiment incorporating a ridge formed by corrugating the heat exchanger tube 30 wall to impress a ridge in the tube inner wall in the shape of a helix 500. As described above, the single helical ridge can be characterized by the elevation angle, α. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s.) The cross-sectional shape of the ridge 510 may be of any shape that produces a protrusion of the inner tube surface into the flow. FIG. 5B shows an expanded view of the ridge of FIG. 5A in the region 511 produced by a corrugation 500 that is circular in cross-section 510 with radius 520 denoted rad 525. Corrugations that are approximately or precisely circular, elliptical, square, rectangular, or trapezoidal are specifically mentioned. Also, corrugations that are approximately or precisely parts of circular, elliptical, square, rectangular, or trapezoidal in cross section are specifically mentioned.

There is a notable advantage to embodiments of heat exchanger tubes with a smooth outer wall: corrugations suffer from material and performance issues, such a surface pitting and deposits (e.g., steam boiler heat exchanger applications) and unwanted hot spot boiling in the second fluid (e.g., hydronic boiler heat exchanger applications. Thus, heat exchanger tube embodiment including a ridge wherein a smooth outer surface is maintained are desirable. FIG. 5C shows an embodiment incorporating a ridge 530 formed by disposing a ridge with semi-circular cross-section 535 on the heat exchanger tube 30 inner wall surface 20 and outer wall surface 10. As described above, the single helical ridge can be characterized by the elevation angle, α. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s.) The cross-sectional shape of the ridge 535 may be of any shape that produces a protrusion of the ridge boundary layer. A semi-circular cross section as illustrated is specifically mentioned. FIG. 5D shows an expanded view of the ridge of FIG. 5C in the region 531 produced by disposing a ridge on the heat exchanger tube 30 inner wall surface 20 that is semi-circular in cross-section 535 with radius 540 denoted rad 545. Ridges that are approximately or precisely circular, elliptical, square, rectangular, or trapezoidal are specifically mentioned. Also, ridges that are approximately or precisely parts of circular, elliptical, square, rectangular, or trapezoidal in cross section are specifically mentioned.

FIG. 5E shows an embodiment incorporating a ridge 550 formed by disposing a ridge with semi-elliptical cross-section 555 on the heat exchanger tube 30 inner wall surface 20. As described above, the single helical ridge can be characterized by the elevation angle, α. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s.) The cross-sectional shape of the ridge may be of any shape that produces a protrusion of the ridge into the flow boundary layer. A semi-elliptical cross section as illustrated is specifically mentioned. FIG. 5F shows an expanded view of the ridge of FIG. 5E in the region 554 produced by disposing a ridge on the heat exchanger tube 30 inner wall surface 20 and outer wall surface 10 that is semi-elliptical in cross-section 555 with major axis 559 denoted as a and minor axis 557 denoted H.

FIG. 5G shows an embodiment incorporating a ridge 560 formed by disposing a ridge with rectangular cross-section 565 on the heat exchanger tube 30 inner wall surface 20 and outer wall surface 10. As described above, the single helical ridge can be characterized by the elevation angle, α. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s.) The cross-sectional shape of the ridge 565 may be of any shape that produces a protrusion of the ridge into the flow boundary layer. A rectangular cross section as illustrated is specifically mentioned. FIG. 5H shows an expanded view of the ridge of FIG. 5G in the region 566 produced by disposing a ridge 560 on the heat exchanger tube 30 inner wall surface 20 that is rectangular in cross-section 565 with base length 569 denoted as w and height 567 denoted H.

FIG. 5I shows an embodiment incorporating a ridge 570 formed by disposing a ridge with trapezoidal cross-section 575 on the heat exchanger tube 30 inner wall surface 20 and outer wall surface 10. As described above, the single helical ridge can be characterized by the elevation angle, α. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s.) The cross-sectional shape of the ridge 575 may be of any shape that produces a protrusion of the ridge into the flow boundary layer. A regular trapezoidal cross section as illustrated is specifically mentioned. FIG. 5J shows an expanded view of the ridge of FIG. 5G in the region 576 produced by disposing a ridge 570 on the heat exchanger tube 30 inner wall surface 20 that is trapezoidal in cross-section 575 with base length 579 denoted as w, peak width 578 denoted as w′ and height 577 denoted H.

To promote clarity in the discussion below, we introduce new nomenclature to describe the fluid dynamic effect near a ridge disposed on the inner surface of a heat exchanger tube. In the disclosure, we refer to vorticity to describe spinning motion of fluid shed from a ridge into the free stream flow—free stream flow being the laminar flow along the longitudinal axis of the heat exchanger tube—as would be seen by an observer located at that point and traveling along with the flow into the free stream. So defined, generally vorticity is undesirable—it creates an obstacle to the free stream flow and increases the pressure drop. Mathematically, the vorticity (as we use the term below) is the curl of the instantaneous flow velocity shed into the free stream (ω=∇v), a vector quantity, and the magnitude of the vorticity is the length of the vorticity vector. We shall use the term swirl (or, equivalently, shear swirl) to describe spinning motion of fluid shed from a ridge into the boundary layer. Generally, shear swell is desirable—it generates mixing in the boundary layer, inhibits the boundary layer from becoming laminar flow and decreases the insulative effects inherent in a laminar flow boundary layer. Mathematically, the shear swirl (as we use the term below) is also the curl of the instantaneous flow velocity, but shed instead into the boundary layer and not into the free stream flow. Moreover, in the disclosure below, shear swirl from the ridges shown in the embodiments have a substantial vector component parallel to the heat exchanger tube wall, spinning or shearing the boundary layer flow against the tendency towards laminar flow.

With this definition, a second aspect of the present disclosure is the surprising discovery that the most effective (e.g., best tradeoff between HFR and DPR) ridge designs do not protrude into the free stream flow region of the heat exchanger tube but, instead, are confined to a fraction of the boundary layer height. FIG. 6A illustrates a longitudinal cross-section of a small region of flow near the inner wall surface 20 of a heat exchanger tube 30 with outer surface 10. Without being bound by theory, the profile 615 of the average flow velocity decreases from its maximum in the free stream 610 until it is zero where the boundary layer meets the heat exchanger tube 30 wall inner surface 20. The height, δ, of this decreasing velocity profile is the boundary layer height 600. When a ridge 620 is disposed on the heat exchanger tube 30 wall inner surface 20, the boundary layer fluid flow is disturbed 630, particularly in the region behind the ridge acting as a bluff body.

It has been surprisingly discovered that effective ridge designs correspond to cases where the ridge height, H, is smaller than or approximately equal to the boundary layer height, δ; that is, H≤δ. In these cases, significant shear swirl is induced in the flow 630 shed downstream of the ridge that has a substantial component parallel to the heat exchanger tube 30 inner wall surface 20 due to the ridge spiral angle, α, not shown in the cross-section. (Equivalently, the spiral angle, β (not shown), or the pitch, P_s(not shown)). This shear swirl persists and contributes to mixing within the flow boundary layer, resulting in a decrease in the insulative effects of a laminar boundary layer that consequently inhibits convective heat transfer from the free stream flow to the heat exchanger tube 30 inner wall and a corresponding reduction in the heat transfer coefficient, h, near the wall. Thus, the ridge 620 disposed on the heat exchanger tube 30 inner wall surface 20 serves to promote efficient heat transfer from the flow free stream through the boundary layer and to the heat exchanger wall 30, resulting in an increase in the local heat transfer coefficient, h_enhanced, and promoting improved bulk heat transfer from a first fluid flowing inside the heat exchanger tube to a second fluid outside the heat exchanger tube. The shed shear swirl and enhanced fluid mixing is approximately confined to the flow boundary layer, does not extend into the free stream and, as a result, results in little or no increase in fluid flow friction or resistance and, hence, contributes little to the pressure drop. The result is different where the ridge protrudes into the free stream flow as illustrated in FIG. 6B. As before, FIG. 6B illustrates a longitudinal cross-section of a small region of flow near the inner wall surface 20 of a heat exchanger tube 30 with outer wall surface 10. Without being bound by theory, the profile 615 of the average flow velocity decreases from its maximum in the free stream 610 until it is zero where the boundary layer meets the heat exchanger tube 30 wall inner surface 20. The height, δ, of this decreasing velocity profile is the boundary layer height 600.

However, in this geometry, the ridge height, H, is significantly greater than the boundary layer height, δ; that is, H>δ. When a ridge 650 is disposed on the heat exchanger tube 30 wall inner surface 20, the boundary layer fluid flow is the flow streamlines 660 are disturbed in the free stream flow, both in front of the bluff body ridge 650 and in the wake 670 of the ridge. It may still be true that the ridge 650 disposed on the heat exchanger tube 30 inner wall surface 20 serves to promote efficient heat transfer from the flow free stream through the boundary layer and to the heat exchanger wall 30, resulting in an increase in the local heat transfer coefficient, h, and promoting improved bulk heat transfer from a first fluid flowing inside the heat exchanger tube to a second fluid outside the heat exchanger tube. However, the improvement in bulk heat transfer comes at a significant cost: the vorticity in the streamlines 670 above the ridge present a flow obstacle to the incoming free stream 660, increasing the pressure drop Δ√{square root over (Pr_AB)} across along the length of the heat exchanger tube as compared to the smooth heat exchanger tube, resulting in an increased DPR. The creation of vorticity that encroaches into the free stream primarily serves to increase the pressure drop, it does not contribute beneficially to the increase in convective heat transfer through the boundary layer to the heat exchanger tube wall.

Thus, a third aspect is the surprising discovery that the geometry of the ridge disposed on the inner surface of the heat exchanger tube—particularly, the ridge height and spiral angle—are key parameters to achieving a desirable tradeoff between increased heat transfer through the boundary layer (HFR>>1) and small increases in pressure drop (DPR≈1) compared to smooth heat exchanger tubes with similar dimensions and material properties. Ridge heights approximately equal to or less than the boundary layer height provide beneficial generation of shear swirl that promotes boundary layer mixing without contributing to pressure drop due to shed vorticity. Empirically, for applications involving combustion gas as a first fluid flowing inside the heat exchanger tube with flow velocities typical of domestic, commercial and industrial applications, ridge heights less than or approximately equal to one millimeter (1 mm) are empirically effective. Ridge heights less than or approximately equal to 0.7 mm is specifically mentioned. Ridge heights less than or approximately equal to 0.6 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.6 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.5 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.45 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.4 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.3 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.2 mm is also specifically mentioned. Ridge heights less than or approximately equal to 0.1 mm is also specifically mentioned.

In practicality, ridge heights are also limited by manufacturing constraints depending upon the manufacturing methods used. For example, cutting, etching, corrugation, embossing, casting, printing are all technologies that can be employed to provide a ridge on the inner surface of the heat exchanger tube. A practical limit for a ridge height of about 0.001 mm also coincides with effective heat transfer effects for commercial and industrial boiler. As a result, ridge heights are approximately 1 mm, or 0.9 mm, or 0.8 mm, or 0.7 mm, or 0.6 mm, or 0.5 mm, or 0.4 mm, or 0.3 mm, or 0.2 mm, or 0.1 mm and 0.001 mm, or 0.01 mm, or 0.02 mm, or 0.03 mm or 0.04 mm, or 0.05 mm. The foregoing upper and lower bounds can be independently combined. The range 0.001 mm to 1 mm is specifically mentioned. The range 0.001 mm to 0.7 mm is specifically mentioned. The range 0.001 mm to 0.4 mm is specifically mentioned.

The spiral angle, β (equivalently, the pitch, P_s, and the elevation angle, α) is key to optimizing the magnitude of the beneficial shear swirl shed behind the ridge. If the spiral angle is too small, the ridge merely presents a bluff body approximately orthogonal to the direction of the flow and, instead of shear swirl behind the ridge, vorticity is created that sheds into the free stream. If the spiral angle is too large, the ridge fails to adequately turn the flow to sharply enough to induce shear swirl in the boundary layer. For a fixed ridge height approximately equal to or less than the boundary layer height, the inventors have surprisingly discovered there is empirically an optimal spiral angle that provides sufficient flow turning to general shear swirl contained primarily within the boundary layer, without shedding significant vorticity. Useful spiral angles have been empirically and by simulation determined to be in the range of greater than or equal to approximately one (1) degree and less than or equal to approximately fifty (50) degrees, depending on the flow inlet velocity and temperature. Spiral angles in the range greater than or equal to approximately ten (10) degrees and less than or equal to approximately forty (40) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately fifteen (15) degrees and less than or equal to approximately thirty-five (35) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately eighteen (18) degrees and less than or equal to approximately thirty (30) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately nineteen (19) degrees and less than or equal to approximately twenty-eight (28) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately nineteen (19) degrees and less than or equal to approximately twenty-seven (27) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately nineteen (19) degrees and less than or equal to approximately twenty-six (26) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately nineteen (19) degrees and less than or equal to approximately twenty-five (25) degrees is also specifically mentioned. Spiral angles in the range greater than or equal to approximately nineteen (20) degrees and less than or equal to approximately twenty-six (25) degrees is also specifically mentioned. A spiral angle of approximately nineteen (19) degrees is specifically mentioned. A spiral angle of approximately twenty (20) degrees is specifically mentioned. A spiral angle of approximately twenty-one (21) degrees is specifically mentioned. A spiral angle of approximately twenty (20) degrees is specifically mentioned. A spiral angle of approximately twenty-two (22) degrees is specifically mentioned. A spiral angle of approximately twenty-three (23) degrees is specifically mentioned. A spiral angle of approximately twenty-four (24) degrees is specifically mentioned. A spiral angle of approximately twenty-five (25) degrees is specifically mentioned. A spiral angle of approximately twenty-six (26) degrees is specifically mentioned. Pitch, P_s, and elevation angle, α, ranges and values corresponding to each spiral angle range and value cited above are also specifically mentioned.

TABLE 1 presents the results of a computational fluid dynamic (CFD) simulation of the effect described above and depicted in FIG. 6A and FIG. 6B. For the simulation results shown a circular cylindrical heat exchanger tube was modeled with a spiral ridge disposed on the inner wall surface and the following parameters: The tube outside diameter, D_O=12.7 millimeters (mm); the tube inside diameter, D_I=10.21 mm; the ridge pitch P_s=9.53 degrees (deg) and the spiral angle β=16.55 deg. The simulated tube length is 16 inches. The inlet flow velocity is 61 meters per second (m/s). The data in TABLE 1A indicates that for a ridge height of 0.43 mm—a value above the boundary layer height—the improvement in thermal heat transfer coefficient compared to a smooth tube with similar geometry; that is, is HFR=3.29. However, the pressure drop increase is substantial; DPR=2.26 indicating that the pressure drop for the heat exchanger tube with a tall spiral ridge is over twice (2.26) the pressure drop of the corresponding smooth tube. Compare these results with the case for a ridge height of 0.21 mm, a value less than the boundary layer height. Here the enhanced heat transfer HFR=2.87, a threefold improvement, but at minimal increase in pressure drop of only DPR=1.20.

TABLE 1A

CFD Simulation Results for Varying Ridge Heights

H = 0.43 mm
H = 0.21 mm
H = 0.15 mm
H = 0.10 mm

DPR
2.26
1.47
1.20
1.13

HFR
3.29
2.97
2.87
2.35

CFD simulations for variations of ridge height and spiral angle (equivalently, pitch) confirm that for a ridge close to or below the height of the boundary layer, a spiral angle exists that is effective at turning the flow to induce swirl in the boundary flow that promotes mixing and increases thermal heat transfer through the boundary layer with minimal additional pressure drop compared to the corresponding smooth tube case. TABLE 1B displays additional CFD simulation cases validating this discovery. All the cases displayed in TABLE 1B correspond to CFD simulations for a straight circular cylindrical heat exchanger tube with a length of 16 inches, 0.5-inch D_O, 10.21 mm D_I, a single semi-circular spiral ridge with the spiral angle and pitch as indicated in the TABLE 1B and an inlet flow velocity of heated combustion gas at 61 meters/second.

CFD simulations have also been used to illuminate the fine details of the flow structure near a ridge disposed on an inner wall surface of a heat exchanger tube. FIG. 7A shows an illustration of a spiral ridge 310 disposed on an inside wall of a heat exchanger tube 30. Also shown are fluid flow trajectories for an ensemble of four distinct initial conditions starting near the ridge 310. The details are shown in a region of the ridge 701 designated FIG. 7B showing the paths of each of the four distinct trajectories leaving the vicinity 717 of the ridge and propagating downstream. The initial conditions for each of the four distinct trajectories begins near the ridge on the downstream side at four different heights measured from the inner heat exchanger wall surface: the initial condition closest to the inner wall surface 710 and three additional trajectories corresponding to the next highest initial condition 712; the third highest initial condition 714; and, finally, the trajectory 716 for a fourth initial condition furthest from the heat exchanger inner wall surface. The initial position 700 of a trajectory in the ensemble is identified.

TABLE 1B

CFD Simulation Results for Varying Ridge Heights

Case
Height,
Pitch, P_s
Angle, β
ΔP

h

#
H (mm)
(mm)
(deg)
(kPa)
DPR
(W/m²K)
HFR

001
1.00
9.53
16.54
5.094
7.69
300.27
3.58

002
0.50
9.53
16.54
1.651
2.49
282.94
3.37

003
0.43
9.53
16.54
1.495
2.26
276.18
3.29

004
0.21
9.53
16.54
0.972
1.47
249.30
2.97

005
0.15
9.53
16.54
0.797
1.20
241.17
2.87

006
0.10
9.53
16.54
0.749
1.13
197.53
2.35

007
0.43
6.55
11.54
1.842
2.78
299.18
3.57

008
0.43
9.53
16.54
1.495
2.26
276.18
3.29

009
0.43
12.66
21.54
1.076
1.62
257.88
3.07

010
0.43
25.40
38.38
0.905
1.37
107.56
1.28

011
0.21
6.55
11.54
1.213
1.83
268.27
3.20

012
0.21
9.53
16.54
0.972
1.47
249.30
2.97

013
0.21
12.65
21.54
0.789
1.19
235.27
2.80

014
0.21
16.02
26.54
0.726
1.10
167.88
2.00

015
0.15
6.55
11.54
1.106
1.67
285.59
3.40

016
0.15
9.53
16.55
0.797
1.20
241.17
2.87

017
0.15
12.65
21.54
0.768
1.16
230.93
2.75

018
0.15
16.02
26.54
0.758
1.14
216.36
2.58

019
0.15
19.69
31.55
0.728
1.10
142.82
1.70

020
0.15
23.77
36.54
0.721
1.09
117.94
1.41

021
0.10
9.53
16.55
0.749
1.13
197.53
2.35

022
0.10
12.65
21.54
0.736
1.11
187.67
2.24

023
0.10
16.02
26.54
0.712
1.08
173.23
2.06

The details of the placement of the ensemble of initial conditions are illustrated in FIG. 7C including the initial position 700 of the third trajectory from the wall inner surface. The ridge 565 height 567, H, is shown for a rectangular ridge of width 569 denoted by w. The ridge height 567, H, was simulated to be smaller than the boundary layer thickness. Each of the four ensemble trajectories 701 is shown in comparison with the ridge height, H, with the two lowest trajectories 710 and 712 selected with initial conditions 717 below the height of the ridge, and two trajectories with initial conditions 714 and 716 selected above the ridge height. These data show that fluid points near or in the boundary layer, stay in the boundary layer. The ridge effectively induces shear swirl in the boundary layer flow—promoting mixing and inhibiting insulative effects—without introducing significant vorticity into the free stream flow that would degrade the pressure drop.

FIG. 7D displays the geometry depicted in the CFD simulation results. For each of the ensemble trajectories (710, 712, 714, 716), the presence of the ridge induces two types of shear swirl in the streamlines: The first is a high-frequency twisting of the streamline superimposed on a long wavelength spiral in the streamline induced by the spiral angle, β, of the ridge. FIG. 7E further illustrates the twisting shear swell 730 induced by the ridge in the streamlines together with its relationship to the instantaneous flow velocity along the streamline 710. From these data it is clear that the local shape of the ridge relative to the bulk flow velocity influences the high-frequency twisting of the streamlines, while the spiral angle, β, (equivalently, the elevation angle, α (not shown), or the pitch, P_s) is important in determining the long-wavelength spiral in the boundary layer flow near the heat exchanger inner wall surface. Both the twisting of the streamlines and the boundary layer spiral motion promotes boundary layer mixing, inhibits the formation of simple insulative laminar boundary layer flow and are controlled by the ridge geometrical parameters.

Shear swirl introduced into the flow streamlines by the ridge at a particular point decay as the streamlines are propagated along the length of the heat exchanger tubes as illustrated in FIG. 7F. Each of the four ensemble streamlines displayed in FIG. 7A through FIG. 7E is shown in a cross-section of the heat exchanger tube 30, from the streamline with initial condition closest to the inner wall 710 to the streamline with initial condition furthest from the wall 710. Superimposed on the heat exchanger tube 30 illustration is a plot of the flow spin vector magnitude 740 (combined vorticity and shear swirl) as the ensemble streamlines propagate down a short section of the heat exchanger tube 30. The magnitude of the combined vorticity and shear swirl (in radians per second) imposed on the streamline at a particular point along the ridge decays over the length of the heat exchanger tube 30. Note that the initial condition ensemble selected for the displayed simulation begins at the heat exchanger tube inlet. Since the free stream flow generates significant vorticity as it enters the tube, the plot shows a substantial contribution due to the inlet vorticity that decays as the ensemble trajectories propagate away for the inlet. The disturbance due to the tube inlet is characteristic of flow near the inlet and decays until the free stream achieves laminar flow. This decay 740 continues along the entire length of the heat exchanger tube 30; however, each section of the ridge downstream away from the inlet is similarly inducing primarily new shear swirl into the boundary layer flow.

A fourth aspect is the surprising discovery that the beneficial increase in thermal heat transfer rate and low increase in pressure drop (compared to a corresponding smooth tube) for a heat exchanger tube with a ridge disposed on the inner tube surface with height equal to or less than approximately the boundary layer height and a spiral angle optimized to contribute shear swirl into the boundary layer persists for heat exchanger tubes of various diameters and lengths. TABLE 1C shows CFD simulation results for configuration of heat exchanger tube lengths from 16 inches to 41 inches.

TABLE 1C

CFD Simulation Results for Varying Heat Exchanger Tube Lengths

Case
Length,
Height,
Pitch, P_s
Angle, β
ΔP

h

#
L (in)
H (mm)
(mm)
(deg)
(kPa)
DPR
(W/m²K)
HFR

024
16
1.00
9.53
16.54
5.094
7.69
300.27
3.58

025
25
1.00
9.53
16.54
5.495
5.98
195.73
2.82

026
41
1.00
9.53
16.54
6.207
4.56
159.31
2.84

027
16
0.43
6.55
11.54
1.842
2.78
299.18
3.57

028
25
0.43
6.55
11.54
2.581
2.81
283.00
4.08

029
41
0.43
6.55
11.54
3.395
2.50
271.21
4.83

030
41
0.10
12.65
21.54
1.485
1.09
169.26
3.01

031
41
0.10
13.64
23.02
1.476
1.09
158.52
2.82

032
41
0.10
13.97
23.54
1.472
1.08
156.29
2.78

033
41
0.10
16.02
26.54
1.426
1.05
105.53
1.88

These results for various tube lengths compare favorably and display the same trends as for the 16-inch tube length simulation shown in TABLE 1B. For a heat exchanger tube of 41 inches, even a spiral ridge height, H, of 0.10 mm—substantially below the boundary layer height—the heat transfer rate for the treated tube is over three times (3.01 times; case number 030) that of a smooth tube of the same dimensions and material, but with almost identical pressure drop (DPR=1.09).

Because the heat transfer thermodynamic mechanism is local to a region of the heat exchanger tube in the boundary layer of the inner tube surface proximal to the ridge, these results persist for heat exchanger tubes of longer length and greater inner tube diameter, D_I. What does vary with tube dimension is the magnitude or contribution of these local thermodynamic effect to the macroscopic benefit of the optimized selection of ridge height, elevation angle and ridge separation since the mass flow rate and the velocity of the fluid or gas flowing through the heat exchanger tube must change from the inlet to the outlet to maintain a constant mass flow rate. As fluid traverses the heat exchanger tube from the inlet to the outlet, the fluid changes (for example, in embodiment where energy is being transferred from the first fluid to the second fluid, the temperature of the first fluid declines) and the fluid velocity changes (for example, in embodiment where energy is being transferred from the first fluid to the second fluid, the velocity of the first fluid declines), also depending upon the motion and homogeneity of the second fluid surrounding the heat exchanger tube. These macroscopic properties affect the local thermodynamics of heat transfer in a local region of the tube boundary layer. Specifically mentioned is that, for example, in embodiment where energy is being transferred from the first fluid to the second fluid, the temperature gradient across the boundary layer is smaller near the heat exchanger tube outlet compared to the tube inlet. Therefore, theory and CFD simulation results predict that the effects of ridge geometry on bulk heat transfer are larger near the inlet of such a heat exchanger tube than near the outlet.

Since the ridge is inducing plentiful shear swirl into the boundary layer flow along its entire length that serves to breakdown the insulative effects that would be otherwise present for a tube with a smooth inner wall surface, additional pressure drop relief may also be achieved by introducing gaps in the ridge structure without prohibitively affecting the benefits of shear swirl creation by the ridge. FIG. 8A illustrates the use of ridge gaps to further relieve the pressure drop penalty. Along the length of the heat exchanger tube 30, gaps 800 are introduced in the ridge that allows free passage of the fluid near the heat exchanger tube 30 inner surface 20. It is convenient to describe the relationship of two adjacent gaps along the ridge spatial curve using the gap separation angle 810, denoted by the angle 820, ψ. FIG. 8B shows the separation angle between adjacent ridge gaps in a cross-section of the heat exchanger tube 30. The ridge gaps separation angle, ψ, may be constant along the ridge spatial curve resulting in a ridge with periodic gaps; however, ridge gap separation angles that are not constant and varies along the ridge spatial curve are also contemplated. In particular, increasing, decreasing and random ridge gap separation intervals are specifically mentioned.

Beside the spacing between adjacent gaps 826, the three-dimensional shape of the ridge gaps can be used by one skilled in the art of heat exchanger tube design to control the boundary layer flow between the passages. FIG. 9A shows a cross-section of a heat exchanger tube 30 together with the inside wall surface 20 and the outside wall surface 10. For two adjacent gaps, 822 and 824, the separation angle can be used to align or misalign the two adjacent gaps with an offset, g, relative to the longitudinal axis. Since a spiral ridge introduces a long-wavelength spiral in the boundary layer flow partially determined by the flow velocity and the spiral angle, β, the offset (equivalently, the ridge gap separation angle) can be adjusted to position the gaps relative to mixing streamlines. Moreover, the shape of the gaps themselves can be used to further induce shear swirl (e.g., off gap corners) to promote additive additional mixing in the boundary or inhibit vorticity (e.g., off top surface edges) where it might shed into the laminar free stream and increase the resulting pressure drop. FIG. 9B through FIG. 9E shows cross-sectional drawings for various shapes of a ridge gap 851 in accordance with embodiments of the present disclosure, not to be taken as limiting the geometries contemplated as part of the present disclosure.

FIG. 9B shows a ridge gap 851 with an approximate rectangular gap shape 850. A rectangular ridge gap is specifically mentioned. The special example of a square ridge gap shape is also specifically mentioned. It is further contemplated that the rectangular ridge gap may not extend for the full height, H, of the ridge. A rectangular ridge gap that extends for the full height, H, of the ridge is specifically mentioned. A rectangular or square ridge gap that extends for only partly into the ridge structure, constituting a notch in the ridge, is also specifically mentioned.

FIG. 9C shows a ridge gap 851 with an approximate semi-elliptical gap shape 860. A semi-elliptical ridge gap is specifically mentioned. The special example of a semi-circular ridge gap shape is also specifically mentioned. It is further contemplated that the semi-elliptical ridge gap may not extend for the full height, H, of the ridge. A semi-elliptical ridge gap that extends for the full height, H, of the ridge is specifically mentioned. A semi-elliptical or semi-circular ridge gap that extends for only partly into the ridge structure, constituting a notch in the ridge, is also specifically mentioned.

FIG. 9D shows a ridge gap 851 with an approximate trapezoidal gap shape 870. A trapezoidal ridge gap is specifically mentioned. The special example of a symmetrical trapezoidal ridge gap shape is also specifically mentioned. It is further contemplated that the trapezoidal ridge gap may not extend for the full height, H, of the ridge. A trapezoidal ridge gap that extends for the full height, H, of the ridge is specifically mentioned. A trapezoidal or symmetrical trapezoidal ridge gap that extends for only partly into the ridge structure, constituting a notch in the ridge, is also specifically mentioned.

FIG. 9E shows a ridge gap 851 with an approximate smooth leading edge gap shape 880. A ridge gap with a fillet edge is specifically mentioned. A ridge gap with a chamfered edge is specifically mentioned.

Since a single ridge with correct geometrical characteristics contained within the flow boundary layer contributes important shear swirl to the boundary layer flow—and thereby, contributes mixing and increasing the heat transfer—without inducing substantial pressure drop, a plurality of ridges disposed on the heat exchanger wall inner surface can have compounding benefits. FIG. 10A shows two spiral ridges disposed on the inner surface of a circular cylindrical heat exchanger tube 30. A first spiral ridge 1000 is disposed on the inner wall surface relative to a second spiral ridge 1010 also disposed on the inner wall surface. Ridge with and without gaps are contemplated. Also, both left and right spiral ridges are contemplated together in the same embodiment. The spiral angles, β₁and β2, may be selected so that the ridges may intersect or run parallel along the length of the heat exchanger tube 30. If the spiral angles, β₁and β₂, are selected for intersection, the tangent line for the first ridge 1020 and the second ridge 1030 will intersect at an angle; it is contemplated that these intersections my be periodic along the length of the heat exchanger tube 30 or aperiodic. It is further contemplated that the intersection angle may be constant along the length of the heat exchanger tube, or non-constant. These variations may be extended to one skilled in the art of heat exchanger tube design to a plurality of ridges. Specifically, two or more ridges disposed on the inner surface of a heat exchanger tube is specifically mentioned. FIG. 10B displays a configuration wherein three ridges 1050, 1051 and 1052 are disposed on the inner surface of the heat exchanger tube in a parallel or nested spiral ridge configuration.

TABLE 1D displays CFD simulation data for heat exchanger tube configurations with gaps as shown in FIG. 9B, and cases with two counter-spiral ridges disposed on the tube inner surface.

TABLE 1D

CFD Simulation Results for Spiral Ridge with Gaps (Cases 034, 035)

and Counter Spiral Ridges (Cases 036 through 040).

Case
Height, H
Pitch, P_s
Angle, β
ΔP

h

#
(mm)
(mm)
(deg)
(kPa)
DPR
(W/m²K)
HFR

034
0.15
12.65
21.54
0.793
1.20
238.81
2.85

035
0.15
12.65
21.54
0.906
1.37
277.21
3.30

036
0.15
12.65
21.54
1.009
1.52
254.35
3.03

037
0.15
16.02
26.54
0.860
1.30
242.96
2.90

038
0.15
12.65
21.54
1.009
1.52
254.35
3.03

039
0.21
12.65
21.54
1.148
1.73
262.76
3.13

040
0.43
12.65
21.54
1.712
2.59
290.01
3.46

Additional embodiments are also specifically mentioned:

EMBODIMENT A: In an embodiment, disclosed is a heat exchanger tube comprising an inlet and an outlet; the heat exchanger tube enclosing and at least partly in contact with flow of a first fluid with the tube inner surface, separated from a second fluid at least partially in contact with the heat exchanger tube outer surface; one or a plurality of ridges disposed on the inner surface of the heat exchanger tube wall; a ridge having a height measured from the inner heat exchanger wall surface of less than approximately 1.0 millimeters.

EMBODIMENT B: In an embodiment, disclosed is a heat exchanger tube with outside diameter, D_O, less than or equal to ten (10) inches and inside diameter, D_I, greater than or equal to one-tenth (0.1) inches, comprising an inlet and an outlet; the heat exchanger tube enclosing and at least partly in contact with flow of a first fluid with the tube inner surface, separated from a second fluid at least partially in contact with the heat exchanger tube outer surface; one or a plurality of ridges disposed on the inner surface of the heat exchanger tube wall; a ridge having a height measured from the inner heat exchanger wall surface of less than or equal to one (1.0) millimeters.

EMBODIMENT C: In an embodiment, disclosed is a heat exchanger tube with outside diameter, D_O, less than or equal to ten (10) inches and inside diameter, D_I, greater than or equal to one-tenth (0.1) inches, comprising an inlet and an outlet; the heat exchanger tube enclosing and at least partly in contact with flow of a first fluid with the tube inner surface, separated from a second fluid at least partially in contact with the heat exchanger tube outer surface; one or a plurality of ridges disposed on the inner surface of the heat exchanger tube wall; a ridge having a height measured from the inner heat exchanger wall surface of less than or equal to approximately one (1.0) millimeters; a ridge having the approximately the shape of a spiral with spiral angle greater than or equal to approximately one (1) degree and less than or equal to approximately fifty (50) degrees.

EMBODIMENT D: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a circle.

EMBODIMENT E: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a semi-circle.

EMBODIMENT F: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a sector of a circle.

EMBODIMENT G: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a rectangle.

EMBODIMENT H: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a square.

EMBODIMENT I: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a trapezoid.

EMBODIMENT I: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately an ellipsoid.

EMBODIMENT J: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a semi-ellipsoid.

EMBODIMENT K: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a sector of an ellipsoid.

EMBODIMENT L: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape of approximately a sector of a polygon.

EMBODIMENT M: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape that includes approximately a chamfer.

EMBODIMENT N: In an embodiment as described in EMBODIMENT C, wherein a ridge has the cross-section shape that includes approximately a fillet.

EMBODIMENT O: In an embodiment as described in EMBODIMENT C, wherein a ridge comprises one or a plurality of gaps wherein the length of a gap is greater than or equal to approximately 0.01 millimeters and less than or equal to approximately ten percent (50%) of the total longitudinal length of a ridge.

EMBODIMENT P: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a semi-circle.

EMBODIMENT Q: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a sector of a circle.

EMBODIMENT R: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a semi-ellipse.

EMBODIMENT S: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a sector of an ellipse.

EMBODIMENT T: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a rectangle.

EMBODIMENT U: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a square.

EMBODIMENT V: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a polygon.

EMBODIMENT V: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a trapezoid.

EMBODIMENT V: In an embodiment as described in EMBODIMENT O, wherein a gap comprises a longitudinal cross-section shape of approximately a triangle.

EMBODIMENT W: In an embodiment as described in EMBODIMENT O, wherein a gap comprises an edge comprising a chamfer.

EMBODIMENT X: In an embodiment as described in EMBODIMENT O, wherein a gap comprises an edge comprising a fillet.

EMBODIMENT Y: In an embodiment, disclosed is a heat exchanger tube with outside diameter, D_O, less than or equal to ten (10) inches and inside diameter, D_I, greater than or equal to one-tenth (0.1) inches, comprising an inlet and an outlet; the heat exchanger tube enclosing and at least partly in contact with flow of a first fluid with the tube inner surface, separated from a second fluid at least partially in contact with the heat exchanger tube outer surface; a plurality of ridges disposed on the inner surface of the heat exchanger tube wall; a ridge having a height measured from the inner heat exchanger wall surface of less than or equal to approximately one (1.0) millimeters; a ridge having the approximately the shape of a spiral with spiral angle greater than or equal to approximately one (1) degree and less than or equal to approximately fifty (50) degrees.

EMBODIMENT Z: In an embodiment as described in EMBODIMENT Y, wherein a first ridge and second ridge disposed on the inner surface of the heat exchanger tube do not intersect along the entire total length, L, of the heat exchanger tube.

EMBODIMENT AA: In an embodiment as described in EMBODIMENT Y, wherein a first ridge and second ridge disposed on the inner surface of the heat exchanger tube intersect along the entire total length, L, of the heat exchanger tube.

EMBODIMENT AB: In an embodiment as described in EMBODIMENT Y, wherein a first ridge and second ridge disposed on the inner surface of the heat exchanger tube intersect along the entire total length, L, of the heat exchanger tube and the angle of intersection in the plane tangent to the inner surface of the heat exchanger tube at the point of intersection, Ψ, is less than or equal to approximately ninety (90) degrees and greater than or equal to approximately zero (0) degrees.

EMBODIMENT AC: In an embodiment as described in EMBODIMENT A, wherein the heat exchanger tube encloses and is at least partly in contact with flow of a first fluid with the tube inner surface, separated from a second fluid that is at least partially in contact with the heat exchanger tube outer surface; one or a plurality of ridges is disposed on the inner surface of the heat exchanger tube wall; the first fluid comprises at least one component is a gas or vapor state.

EMBODIMENT AD: In an embodiment as described in EMBODIMENT AC, wherein the first fluid comprises at least one component is a gas or vapor state formed by combustion.

EMBODIMENT AE: In an embodiment as described in EMBODIMENT AC, wherein the first fluid comprises at least one component is a gas or vapor state formed by heating a fluid to at least a partially gaseous state.

EMBODIMENT AF: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises water.

EMBODIMENT AG: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises steam.

EMBODIMENT AH: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises cooking oil.

EMBODIMENT AI: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises a petroleum hydrocarbon.

EMBODIMENT AJ: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises an organic chemical.

EMBODIMENT AK: In an embodiment as described in EMBODIMENT AC, wherein the second fluid comprises an inorganic chemical.

In any embodiment disclosed or an equivalent, a ridge comprises a separate element from the heat exchanger tube and disposed on the inner surface of the heat exchanger tube and secured by friction between at least a portion of the surface of ridge element and at least a portion of the inner surface of the heat exchanger tube.

In any embodiment disclosed or an equivalent, a ridge comprises a separate element from the heat exchanger tube and disposed on the inner surface of the heat exchanger tube and secured by weld between at least a portion of the surface of ridge element and at least a portion of the inner surface of the heat exchanger tube.

In any embodiment disclosed or an equivalent, a ridge is disposed on at least a portion of the inner surface of the heat exchanger surface by removing (equivalently, cutting, extracting, impressing, extruding) a channel (equivalently, groove, relief) of material from the inner surface of the heat exchanger tube leaving a ridge structure materially continuous with at least a portion of the heat exchanger wall.

The terms “a” and “an” do not denote a limitation of quantity, but rather denote the presence of at least one of the referenced item. The term “or” means “and/or” unless clearly indicated otherwise by context. Reference throughout the specification to “an embodiment”, “another embodiment”, “some embodiments”, and so forth, means that a particular element (e.g., feature, structure, step, or characteristic) described in connection with the embodiment is included in at least one embodiment described herein, and may or may not be present in other embodiments. In addition, it is to be understood that the described elements may be combined in any suitable manner in the various embodiments. “Optional” or “optionally” means that the subsequently described event or circumstance may or may not occur, and that the description includes instances where the event occurs and instances where it does not. The terms “first,” “second,” and the like, “primary,” “secondary,” and the like, as used herein do not denote any order, quantity, or importance, but rather are used to distinguish one element from another. The terms “front”, “back”, “bottom”, and/or “top” are used herein, unless otherwise noted, merely for convenience of description, and are not limited to any one position or spatial orientation.

The endpoints of all ranges directed to the same component or property are inclusive of the endpoints, are independently combinable, and include all intermediate points. For example, ranges of “up to 25 N/m, or more specifically 5 to 20 N/m” are inclusive of the endpoints and all intermediate values of the ranges of “5 to 25 N/m,” such as 10 to 23 N/m.

Unless defined otherwise, technical and scientific terms used herein have the same meaning as is commonly understood by one of skill in the art to which this invention belongs.

All cited patents, patent applications, and other references are incorporated herein by reference in their entirety. However, if a term in the present application contradicts or conflicts with a term in the incorporated reference, the term from the present application takes precedence over the conflicting term from the incorporated reference.

The following TABLE 2 summarized the nomenclature used to describe the embodiments disclosed:

TABLE 2

SELECTED NOMENCLATURE

D_I
Tube inside diameter

D_O
Tube outside diameter

DPR
Ratio of pressure drop of an enhanced heat exchanger tube

and the pressure drop for an equivalent smooth heat exchanger

tube

HFR
Ratio of the thermal transfer coefficient of an enhanced heat

exchanger tube and the thermal transfer coefficient for an

equivalent smooth heat exchanger tube

L
Tube shape curve length

L_O
Tube overall length

P_rA
Heat exchanger tube inlet pressure

P_rB
Heat exchanger tube outlet pressure

ΔPr_AB
Heat exchanger tube pressure drop (inlet to outlet)

ΔP_smooth
Smooth heat exchanger tube pressure drop (inlet to outlet)

ΔP_enhanced
Heat exchanger tube pressure drop (inlet to outlet) with

treatment (e.g., ridge)

r_R, ω_R, z_R
Cylindrical coordinates along the heat exchanger tube

τ, ρ, λ
Ridge coordinate frame

R
Ridge shape curve

S
Heat exchanger tube shape curve

THK
Tube material wall thickness

x, y, z
Cartesian coordinates

HEAT EXCHANGER TUBE WITH HIGH HEAT TRANSFER AND LOW PRESSURE DROP

Information

Publication Number

Date Filed

Date Published

Inventors

Original Assignees

CPC

International Classifications

Abstract

Description

Claims

REFERENCE TO RELATED APPLICATIONS

Provisional Applications (1)