Patent Application
20200182561
This disclosure relates to tube-in-tube heat exchangers, particularly, tube-in-tube heat exchangers having spiral corrugations.
Tube-in-tube heat exchangers are used to transfer heat from one fluid to another. Tube-in-tube heat exchangers usually consist of two tubes (i.e., an inner tube and an outer tube), in which the outermost diameter of the inner tube is smaller than the innermost diameter of the outer tube so that the inner tube fits completely inside the outer tube. A first fluid flows inside the inner tube, while a second fluid flows in the annulus formed between the exterior surface of the inner tube and the inner surface of the outer tube. The two fluids flow and exchange heat conducted through the wall of the inner tube, without direct contact between the two fluids. Fluids can be in any flowable form (e.g., liquid, gas, two-phase, with and without entrained solids and the like), and the fluids may flow through a heat exchanger in a co-flow (i.e., same direction) or counter-flow (i.e., opposite direction).
There are several variables that impact the rate of heat transfer and the resulting pressure loss. For a simple “smooth” tube-in-tube heat exchanger, the inside diameter (Di_i) of the inner tube establishes the mass flux, velocity, convective heat transfer coefficient, and pressure loss per unit length for the first fluid flowing inside the inner tube. The spatial difference between the inside diameter of the outer tube (Do_i) and the outside diameter of the inner tube (Di_o) defines an annular gap in which the second fluid flows. The resulting hydraulic diameter (Do_i−Di_o) establishes the mass flux, velocity, convective heat transfer coefficient, and pressure loss per unit length for the second fluid flowing in the annulus.
The basic “smooth” or “straight-sided” tube-in-tube geometry often has limitations that prevent an optimum heat exchanger design. For example, standard tubes are only available in a limited number of diameter and wall thickness combinations, limiting the available options for the key geometric variables (Di_o and Di_i). Often, the characteristics (type of fluid, thermophysical properties and the like) of the two fluids are quite different and the respective flow rates and allowable pressure loss may differ as well. If one fluid is a gas that has a low pressure loss requirement, it will almost necessarily have to flow through the inner tube that is large in diameter. Since the flow area of the annulus (and resulting fluid velocity) is a function of the diameter squared, the resulting annular flow area may be larger than desired (causing low fluid velocities and low heat transfer rates) if the other fluid is a liquid at a low flow rate. Even if the inside diameter of the outer tube (Do_i) and the outside diameter of the inner tube (Di_o) can be closely matched so that annulus side velocity increases, the pressure loss penalty for a viscous fluid flowing through a very narrow gap is often very high. The result is the inability to arrive at a desired heat exchanger geometry where the size of the heat exchanger is constrained by the annulus-side fluid heat transfer rate or resulting pressure loss.
Additionally, it is often desirable to coil the resulting length of concentric tubes so that the heat exchanger fits within a smaller package size. However, when two tube-in-tube smooth tubes are coiled, the inner tube does not remain centered (concentric) with respect to the outer tube. Instead, its center axis shifts towards the inside of the coil until the outside diameter of the inner tube rests against the inside diameter of the outer tube. While this may not impact the flow, heat transfer and pressure loss per unit length of the fluid flowing inside the inner tube, the resulting non-concentric annular space negatively impacts the flow and heat transfer characteristics of the fluid flowing in the annular space. The flow in the annulus is no longer evenly distributed around the outside diameter of the inner tube, reducing the effective heat transfer area and increasing the length of the heat exchanger (and resulting pressure loss for both fluids).
Additionally, the hydraulic diameter for the resulting non-concentric annular flow area is no longer simply defined as Do_i−Di_o. The resulting crescent-shaped annual gap that varies along the radial axis is a function of the flow area and the wetted perimeter (4*A/WP), which normally results in a larger hydraulic diameter compared to a concentric annulus and, therefore, lower heat transfer rates.
Various methods have been provided to keep the inner tube concentric after coiling, including helically wrapping the inner tube with a wire (e.g., US 2015/0159957) or forming dimples in the outer tube (e.g., U.S. Pat. No. 2,259,433). The helically wrapped wire can introduce turbulence, mixing, and a spiral flow pattern in the annulus, which are often desirable features that can increase the annulus-side heat transfer coefficient. However, the annulus side pressure loss per unit length often increases dramatically and the cost of wrapping and attaching (normally welding or brazing) the wire to the inner tube can be very expensive. Additionally, while dimples are effective at keeping the inner tube concentric and in some cases slightly increasing the annulus side heat transfer coefficient due to increased mixing, dimples are not suitable for a heat exchanger where the annulus side constrains the heat exchanger design due to heat transfer or pressure loss.
Flow in smooth tubes most often results in low heat transfer performance due to the lack of turbulence, mixing, and the buildup of a thick boundary layer on the smooth wall. Numerous methods and geometries to enhance heat transfer performance (normally with a corresponding pressure loss penalty) in tube-in-tube heat exchangers have been developed. Many of those methods utilize fins or fin-like structures located in the annular space between the two tubes and/or inside the inner tube, or inserts that introduce turbulence (e.g., U.S. Pat. Nos. 2,692,763, 4,286,653, 6,098,704, US 2010/00193168, US 2015/0159957).
When a twisted or fluted tube is used as the inner tube of a concentric tube-in-tube heat exchanger, both the tube-side and annulus-side heat transfer rate is increased (e.g., US 2009/0159248). However, the geometry of fluted tubes (flute height, pitch, number of flutes) is constrained by fabrication and material property limitations. Due to such geometrical constraints, it is often not possible to fabricate a concentric tube-in-tube heat exchanger with an inner fluted tube that provides the desired balance between heat transfer enhancement and pressure loss for both fluids.
Accordingly, there remains a need for a heat exchanger that can maintain a concentric tube-in-tube arrangement, enhance heat transfer and minimize pressure loss for both fluids.
I provide a tube-in-tube heat exchanger comprising an inner tube and an outer tube. The outer tube has at least one spiral corrugation and an inner surface of the corrugation is in contact with or in proximity to the exterior surface of the inner tube to provide a spiral annular channel in an annular space between the inner tube and outer tube.
It will be appreciated that the following description is intended to refer to specific examples of structure selected for illustration in the drawings and is not intended to define or limit this disclosure other than in the appended claims.
This disclosure provides a tube-in-tube heat exchanger where the heat transfer performance and pressure loss of two fluids flowing through the heat exchanger can be controlled by specifying the diameter of the inner and outer tubes and the corrugation pitch of the outer tube. This disclosure also provides a simple and inexpensive method of maintaining concentricity of the inner tube if the tube-in-tube heat exchanger is bent or coiled. The resulting heat exchanger geometry is simple and inexpensive to manufacture, provides for a heat exchanger that can be controlled for the desired performance, and is easily adaptable to a wide range of applications by setting the desired diameter and pitch variables.
Referring to the drawings,
As shown in
The arrangement of the inner surface 106 of the spiral corrugation 108 of the outer tube 104 in contact with or in proximity to the exterior surface of the inner tube 102 forms a spiral annular channel 110 in the annular space between the inner tube 102 and outer tube 104. Fluid in the annular channel 110 can flow in a spiral path around the outside of the inner tube 102 in the annular gap formed between adjacent portions of the corrugation 108, the inner surface of the outer tube 104, and the exterior surface of the inner tube 102. The spiral flow direction of fluid in the annular gap is depicted by the bold arrows in
Length “g” defines the height of the annular channel 110. Length “g” is the distance between the inner surface of the outer tube 104 and the exterior surface of the inner tube 102. A suitable length “g” may be at least 0.01 inches. A preferred range for “g” is about 0.01 inches to about 0.5 inches or, more preferably, about 0.01 inches to about 0.1 inches.
P* is the actual width of the flow area based on the tube diameter and corrugation pitch (P). The perpendicular length (P*) between the adjacent portions of the corrugation 108 is calculated based on the corrugation pitch (P) and the outside diameter of the inner tube 102 (Di_o):
P*=P×sin(Θ)
Θ=arctan(Di_o×3.14/P)=helix angle of corrugation
If there is more than one corrugation (i.e., more than one spiral groove), P* is calculated based on the above equation and divided by the number of corrugations.
The flow velocity and dimensions of the annular channel 110 are key factors determining the heat transfer performance and pressure loss for the fluid flowing in the annular channel 110. Therefore, by varying the outside diameter of the inner tube 102 (Di_o), the inside diameter of the outer tube 104 (Do_i), the length between corrugations (pitch) and ensuring the inner surface 106 of the corrugation 108 is in contact with or is in proximity to the exterior surface of the inner tube 102, the heat transfer performance and pressure loss for the fluid flowing in the annular space can be precisely controlled.
The corrugations 108 in the outer tube 104 also provide a centering mechanism for the inner tube 102 if the concentric tubes (102 and 104) are coiled so that desired heat transfer performance is maintained.
The geometry of the inner tube 102 can be varied depending upon the application and the desired heat exchanger performance. The inner tube 102 can be a simple smooth tube that provides no heat transfer enhancement, or a corrugated tube to provide heat transfer enhancement to the fluid flowing in the inner tube 102.
If a corrugated tube is used for both the inner tube 102 and the outer tube 104, the corrugation pitch (P) may be the same or different for the tubes. A suitable pitch for the corrugations of either the inner tube 102 or the outer tube 104 may be from about 0.25 to about 5, more preferably, from about 0.3 to about 3. The corrugation depth may be the same or different for the two tubes. The direction of spiral may be the same or different (one clockwise and the other counter-clockwise for example).
Other forms of heat transfer enhancement can be used for the fluid flowing inside the inner tube 102, totally independent of the annular flow geometry in the annular channel 110, including internally formed fins, ridges, inserts or the like that promote turbulence and heat transfer enhancement.
The number of corrugation grooves in the outer tube or inner tube (if corrugated) can be one or more than one. If there are two corrugation grooves, the annular flow will be divided into two adjacent spiral flow paths defined by the gap between outer tube and inner tube and the distance between corrugations. If there are three corrugation grooves, the annular flow will be divided into three flow paths, and so on. Preferably, the number of corrugation grooves is between 1 and 4.
The material of the tubes is not particularly limited and can be selected based on the desired application of the heat exchanger. Suitable materials include metal, plastic, elastomer and the like. The tubes of a heat exchanger may be made from the same material or from different materials.
The tube-in-tube heat exchangers are especially useful for heat exchangers in which the two fluids flowing through the heat exchanger are of different forms (liquid or gas, for example), have different flow rates, have different pressure loss specifications, and/or a phase change is occurring for one of the fluids (evaporation or condensation). This is due to the fact that the heat transfer rates can be balanced between the two fluids by choosing the tube diameters and corrugation pitch. Preferably, the fluid that must have a low pressure loss through the heat exchanger is selected to flow inside the inner tube and the fluid that can withstand a higher pressure loss through the heat exchanger is selected to flow through the annular channel, where the velocity can be controlled by the chosen geometry to be higher. Preferably, the pressure loss inside the tube (DELTAP_t) is less than the pressure loss in the annulus (DELTAP_a). In preferred examples, a ratio of (DELTAP_t):(DELTAP_a) is about 0.9 or less, more preferably, about 0.8 or less or, even more preferably, about 0.5 or less.
A heat exchanger application where the tube-in-tube heat exchangers are especially useful is a refrigerant sub-cooler (RSC) for a heat pump. In an RSC, one fluid is the cold, low-pressure refrigerant vapor exiting the evaporator which may still contain some amount of unevaporated liquid. The other fluid is hot high-pressure refrigerant liquid exiting the condenser. In the RSC, the hot high-pressure liquid is cooled to a temperature below its saturation temperature by the cold low pressure refrigerant 2-phase mixture.
The heat transfer coefficient for the low pressure 2-phase mixture will be very high without enhancement due to evaporation of the remaining liquid as it passes though the RSC. However, it is desirable to keep the pressure loss of the low pressure 2-phase refrigerant as low as possible to ensure maximum efficiency of the heat pump cycle. Therefore, it is desirable for the low-pressure refrigerant flow to be inside the inner tube 102 of sufficient size so the velocity is low and pressure loss is minimized. Since the allowable pressure loss of the high-pressure refrigerant liquid flowing in the annulus is much higher than the low-pressure 2-phase refrigerant flowing inside the inner tube, the corrugated tube-in-tube geometry allows the designer to take advantage of this design condition and a shorter (less expensive) heat exchanger results.
In the RSC, the heat transfer coefficient for the high-pressure refrigerant liquid will be much lower given that no phase-change process occurs. Since the fluid is a liquid and, therefore, of much higher density than the low pressure vapor, it can flow through a much smaller flow area without a large pressure loss penalty. Also, since the high-pressure refrigerant liquid will be expanded to the low-side pressure before entering the evaporator after exiting the RSC, the high-pressure liquid refrigerant can have a high-pressure loss through the RSC without impacting the efficiency of the heat pump.
Therefore, the diameter of the inner tube can be selected to provide the desired (low) pressure loss for the low-pressure refrigerant. The outside diameter of the outer tube and the pitch of the corrugations in the outer tube are then selected so that the velocity of the high-pressure refrigerant liquid is high and the heat transfer performance of the high-pressure liquid flowing spirally in the annulus closely matches that of the low pressure 2-phase fluid. The high velocity of the high-pressure refrigerant liquid can cause a high-pressure loss through the RSC but, in this application, that is acceptable.
Another heat exchanger application where the tube-in-tube heat exchangers are especially useful is the solution heat exchanger (SHX) in an absorption heat pump. In the SHX, thermal energy is transferred and recuperated from the hot “weak solution (WS)” exiting the Desorber (solution with low refrigerant content) to the cool “strong solution (SS)” entering the Desorber (solution with high refrigerant content). Although both fluids are at the high-side pressure, the SS has a higher flow rate than the WS and it is preferred to have a low pressure loss through the SHX so that additional back pressure is not put on the solution pump (increasing its required power and reducing reliability). Since the WS will be expanded to the low-side pressure after exiting the SHX, a high pressure loss for the WS in the SHX is acceptable.
Therefore, the diameter of the inner tube can be selected to provide the desired pressure loss for the SS flowing inside the inner tube. Since the SS is a liquid, use of a corrugated inner tube or an insert inside the inner to tube to induce turbulence is desirable. The outside diameter of the outer tube and the pitch of the corrugations in the outer tube are then selected so that the heat transfer performance of the WS flowing in the annulus matches or is higher than the SS flowing inside the enhanced inner tube. The resulting high velocity and, therefore, high-pressure loss of the WS is acceptable because a high-pressure loss for the WS flowing through the SHX does not impact the efficiency of the heat pump.
Many other heat exchanger applications where the two fluids are mismatched in flow rate, phase or allowable pressure loss can benefit from the heat exchanger.
Solution heat exchangers (SHX) for a nominal 80,000 Btu/hr gas absorption heat pump (GAHP) were designed (sized) using a corrugated inner tube inserted into a corrugated outer tube where the inner surface of the corrugations of the outer tube were in contact with or in proximity to the exterior surface of the inner tube. The strong solution was selected to flow inside the inner tube and the weak solution was selected to flow inside the annular channel.
For Examples 1-17, the outside diameter and wall thickness of both the inner and outer tubes was varied, along with the pitch of the corrugations in the outer tube. The calculated heat exchanger lengths ranged from 18 to 60 feet. The geometries of the heat exchangers of Examples 1-17 are shown in Table 1.
By selecting different tube diameter and wall thickness combinations, different annular gap (g) dimensions can be obtained and these gaps are short enough to allow high velocities (and therefore high heat transfer coefficients) for the weak solution flowing in the annulus. The corrugation pitch was also varied to create smaller or larger flow areas (and therefore velocities) in the annulus.
The pressure loss inside the tube (DELTAP_t), the pressure loss in the annulus (DELTAP_a), and overall heat transfer coefficient (U_o) for each example are shown in Table 1.
Solution heat exchangers (SHX) for a nominal 80,000 Btu/hr gas absorption heat pump (GAHP) were designed (sized) using an inner fluted (or twisted) tube inserted into a smooth outer tube, where the inside diameter of the outer tube is in contact with or very close proximity to the outside diameter of the inner fluted tube (tips of the flutes). The dimensions of the heat exchangers of Comparative Examples 1-8 are shown in Table 2. These tubes were selected based on having the shortest flute heights (e) believed possible based on the data base of manufacturable fluted tube geometries published in “A Manual for Heat Exchanger Design Using Spirally Fluted Tubes,” under Gas Research Institute contract #5092-243-2357. The shorter flute heights (g) provide the highest velocity and heat transfer coefficients for the weak solution flowing in the annulus. The heat exchanger lengths ranged from 142 to 252 feet as shown in Table 2, representing a significant increase in heat exchanger size compared to the corrugated tubes of Examples 1-17.
In the SHX, hot weak solution (a liquid low in refrigerant concentration) transfers heat to a strong solution (a liquid of high refrigerant concentration). The flow rate of the strong solution is higher than the weak solution. The allowable pressure loss of the strong solution was kept very low (to prevent back-pressure on the pump), while the allowable pressure loss of the weak solution can be much higher. For these reasons, the strong solution was designated to flow inside the inner tube and the weak solution flows in the annulus between the inner fluted tube and the outer plain tube.
The pressure loss inside the tube (DELTAP_t), the pressure loss in the annulus (DELTAP_a), and overall heat transfer coefficient (U_o) were calculated. Results of SHX designs using the eight selected fluted tube geometries are shown in Table 2.
For all Comparative Examples 1-8, the heat exchanger size is limited by the heat transfer coefficient in the annulus (alpha_a), which is much lower than inside the tube (alpha_t). This result can be seen by the overall heat transfer coefficient (U_o) which is very close to alpha_a. Also, for most of Comparative Examples 1-8, the pressure loss inside the tube (DELTAP_t) is higher than the pressure loss in the annulus (DELTAP_a), which is the opposite of what is desired.
For the corrugated arrangements of Examples 1-17, the annulus side heat transfer coefficients (alpha_a) and pressure losses (DELTAP_a) are much higher than for the fluted tube designs of Comparative Examples 1-8, resulting in a significantly higher overall heat transfer coefficient (U_o) and much shorter heat exchanger lengths. Also, the pressure loss in the annulus (DELTAP_a) is higher than the inside tube (DELTAP_t). Since the allowable pressure loss of the weak solution flowing in the annulus is much higher than the strong solution in the tube, the corrugated tube-in-tube geometry allows the designer to take advantage of this design condition and a shorter (less expensive) heat exchanger results.
Two heat exchangers were fabricated and tested in a GAHP prototype using the tube and heat exchanger dimensions shown in Example 12. The heat exchanger of Example 18 was 20 foot long (4 foot shorter than design) and provided an effectiveness of 0.92 (0.97 design target). The heat exchanger of Example 19 was 30 foot long and provided an effectiveness of 0.98.
Although the apparatus and methods have been described in connection with specific forms thereof, it will be appreciated that a wide variety of equivalents may be substituted for the specified elements described herein without departing from the spirit and scope of this disclosure as described in the appended claims.
This invention was made with government support under Grant DE-EE0006116 awarded by the Department of Energy. The government has certain rights in the invention.
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/US2017/041041 | 7/7/2017 | WO | 00 |
| Number | Date | Country | |
|---|---|---|---|
| 62361580 | Jul 2016 | US |
