1. Field of the Invention
The subject invention relates to heat exchangers, and more specifically to an evaporator, that utilizes flat tubes having a plurality of flow passages extending therethrough.
2. Description of the Related Art
Evaporators for automobile heating, ventilation and air conditioning (HVAC) systems are well known in the art as described in the U.S. Pat. Nos. 4,470,455 and 4,535,839. Such evaporators typically include a core formed by a plurality of tubes between which fins are disposed for permitting ambient air to flow across the exterior of the tubes. The tubes are in fluid communication with spaced tanks to allow refrigerant—working fluid of the system capable of undergoing transformation from liquid to vapor and vice versa—to flow from one tank to the other through the tubes. This permits heat exchange between the refrigerant and the ambient air as the refrigerant flows through the tubes.
Various evaporator tubes exist in the art. For example, a laminated tube is fabricated by joining a pair of embossed plates together to create interior sidewalls that define a channel through which the refrigerant flows. The hydraulic diameter of such a channel is typically determined by multiplying the cross sectional area of the channel by four and dividing that result by the wetted perimeter of the channel. The relatively small hydraulic diameter of the channel and the embossed surfaces of the conjoined plates produce a relatively high convective heat transfer coefficient for the refrigerant flowing through the tube. Despite this advantage, laminated tubes have certain drawbacks. For example, the embossed patterns on the surfaces of the plates make it difficult for the fins to bond to the surfaces. Furthermore, the plates are expensive to fabricate and result in tubes that can be subjected to relatively low refrigerant side pressure.
Certain flat tubes with a plurality of non-circular flow passages fabricated by using extrusion techniques do exist, which are designed to address the drawbacks associated with the laminated tube evaporator as described in the U.S patents bearing the U.S. Pat. Nos. 5,318,114; 6,161,616 and 6,449,979. However, none of these patents deal with the optimal dimensions of the circular or noncircular refrigerant flow passages within the extruded flat tubes nor do they deal with the optimal number of tubes in each pass of a multi-pass evaporator. The present invention is directed at high performance flat tube evaporators with enhanced refrigerant side passages of optimal dimensions and optimal number of tubes in each pass of a multi-pass evaporator.
The dominant heat transfer mechanism within the prior art evaporators is forced convection boiling, which is driven by the flow of the refrigerant through the flow channels. Forced convection boiling typically includes four stages. The first stage, or bubbly flow regime, is that in which the vapor mass fraction of the refrigerant is very low. In the second stage, or slug flow regime, the vapor volume fraction increases and individual bubbles begin to agglomerate to form plugs, or slugs, of vapor that move through the tube. The third stage, or annular flow regime, occurs when the interior walls of the tube are covered with a thin film of liquid refrigerant through which heat is absorbed. The mist flow regime is the final stage. During this stage, there is a sharp reduction in the boiling heat transfer coefficient of the refrigerant within the tube. Throughout all four stages, a nucleate boiling regime exists in selected areas of the tube, which results in quasi pool boiling of the refrigerant in those areas. However, the prior art tubes are not designed to ensure that such boiling optimizes the amount of heat transferred through the tube.
Accordingly, the subject invention overcomes the limitations of the related art by providing a heat exchanger of the type in which a cross-flow of a fluid is directed in an upstream to downstream direction on the external surface of the heat exchanger to induce a transfer of thermal energy between the external fluid and a refrigerant circulating within the heat exchanger. The heat exchanger includes a pair of spaced tanks. A plurality of heat exchange tubes extends between the tanks in fluid communication therewith. At least one of the tubes includes a plurality of flow passages whose interior sidewalls define at least one corner having an included angle of less than ninety degrees to promote intense quasi pool boiling. Reducing the included angle of the corner increases the volume of the liquid refrigerant drawn into the corner by surface tension. This not only enhances nucleate boiling, but also creates secondary flow patterns normal to the primary flow of the refrigerant along the longitudinal axis of the passage defined by the interior sidewalls. An increase in the secondary flow causes a corresponding increase in turbulence within the passage, which further enhances quasi pool boiling and increases the rate of heat transfer through the tube.
Other advantages of the present invention will be readily appreciated as the same becomes better understood by reference to the following detailed description when considered in connection with the accompanying drawings wherein:
Referring to the Figures, wherein like numerals indicate like or corresponding parts throughout the several views, a heat exchanger is generally shown at 40 in
The heat exchanger 40 has an unfolded core design and includes a pair of spaced tanks 42 comprising a plurality of flow separators 68, shown clearly in
The heat exchanger 40 also includes spaced upper and lower reinforcing plates 60 between which the tubes 44 and fins 58 are positioned. The reinforcing plates 60 extend parallel to the tubes 44 and interconnect the tanks 42 to form the heat exchanger core. A selected one of the tanks 42 includes an inlet tube 62 and an outlet tube 64. In
Referring to
In
The number of flow separators 68 is always one less than the number of desired flow passes. When there are no flow separators 68 in the tanks 42, the refrigerant enters the heat exchanger 40 through the inlet tube 62 located in one tank 42 and exits through the outlet tube 62 located in the opposing tank 42. Such a heat exchanger can be characterized as a single-pass heat exchanger since in such a heat exchanger the refrigerant makes a single pass across the external fluid.
Referring now to
As is shown in
In
Referring now to
Referring now to
Although the flow passage 48 may have any shape, the flow passage 48 shown in
Referring back to
The turbulent flow through a circular passage is predominantly unidirectional with only turbulent flow characteristics. On the other hand, the turbulent flow through a noncircular passage, like 48 with sharp corners 50, is bidirectional possessing both turbulent and laminar flow characteristics. The turbulently flowing refrigerant is drawn into the corner regions by the surface tension effect, which gives rise to a non-zero transverse velocity component normal to the interior sidewalls 46. This velocity component, significantly smaller than the turbulent axial velocity component, is laminar in characteristic due to quasi-stagnant nature of the liquid pool formed in the corner region and depends solely on the shape of flow passage 48. Thus springs into existence a coexisting laminar flow within a noncircular passage 48 with sharp corners and turbulently flowing fluid through the flow passage 48. It is found that the coexistence of the laminar flow is particularly predominant when the radius of the corner 50 is small with the included angle “θ” less than or equal to thirty degrees.
Referring now to
The secondary flow does not exist in a circular flow passage with uniformly varying passage wall curvature. Presence of a surface discontinuity in the passage wall is a necessary condition for the existence of a secondary flow in a noncircular flow passage. The surface discontinuity need not be sharp like a knife-edge. It can be a relatively mild discontinuity with non-uniformly varying wall curvature as in an elliptical flow passage. It is only in the limit when an elliptical passage degenerates into a circular passage with uniformly varying wall curvature that the secondary flow disappears.
The mean velocity of the primary flow as well as that of the secondary flow 80 depends solely on the coordinates of the cross section of the flow passage 48. The mean velocity of the secondary flow 80 is approximately 1% to 2% of the mean velocity of the primary flow. Notwithstanding the low magnitude of the secondary flow mean velocity, it exerts a measurable effect in increasing the friction factor coefficient and the heat transfer coefficient for the flow passage. Both of these coefficients are approximately 10% greater in the corners 50 dominated by the secondary flow 80 than in the areas of the tube 44 dominated by the primary flow.
Referring now to
The heat transfer rate through the tubes 44, 144 with flow passages set forth in
When flowing through the tubes in a condenser, the refrigerant changes from a single-phase vapor to a two-phase mixture of saturated liquid. In this case since a higher percentage of the refrigerant in the first pass is in the vapor phase as compared to the liquid phase, the density of the refrigerant in the first pass is smaller than the density of the refrigerant in the last pass. Thus, the number of tubes to be included in each flow pass must progressively decrease from the first to the last pass in a condenser.
Table 1 sets forth the fractions of the optimal number of tubes to be apportioned in each pass of an evaporator. Row 1 of Table 1 indicates the number of flow passes ranging from 1 to 10. Column 1 gives the fraction of the tubes to be apportioned to the single pass of the one-pass evaporator. Clearly the number of tubes that can be assigned to the single pass of a one-pass evaporator equals the total number of tubes in the evaporator. Hence the ratio of the number of tubes in the one pass to the total number of tubes in the evaporator is 1. Column 2 indicates the optimal number of tubes that can be assigned to a two-pass evaporator. The tabular results show that the optimal ratio of the number of tubes in pass P1 to the total number of tubes in the two-pass evaporator is 0.3981 while the optimal ratio of the number of tubes in pass P2 to the total number of tubes in the two-pass evaporator is 0.6019. Similarly, columns 3 through 10 indicate the optimal ratios of the number of tubes in each pass of a three-pass through a ten-pass evaporator.
The results of Table 1 are also represented in the form of a bar chart in
To illustrate the manner in which Table 1 is used, assume that a single evaporator core, as shown in
Number of tubes in first pass=60×0.2153=12.9≅13
Number of tubes in second pass=60×0.2384=14.3≅14
Number of tubes in third pass=60×0.2616=15.7≅16
Number of tubes in fourth pass=60×0.2847=17.1≅17.
Referring now to
wherein,
wherein
In order to calculate the optimal hydraulic diameter “d” of a noncircular passage, the optimal hydraulic diameter “do” of a baseline circular passage must first be determined using Equation (1) in conjunction with the graph set forth in
By way of an example, suppose that a refrigerant flows through an evaporator core in the form of a mixture of saturated liquid and vapor. In order to determine the properties of such a mixture, the properties of the saturated liquid and saturated vapor are required. The refrigerant quality “χ”, which is the vapor mass fraction as a weighting factor for the properties of the mixture, is also required. Although any suitable refrigerant may be utilized with the subject invention, by way of non-limiting example, R-134a is utilized in the examples set forth herein assuming that refrigerant R-134a is flowing through the evaporator core at a temperature of 50° F. and has an average refrigerant quality “χ”=0.7. The transport properties for R-134a refrigerant at a temperature of 50° F. are set forth in Table 2. Throughout Table 2, the subscript “f” denotes the saturated liquid and the subscript “g” denotes the saturated vapor.
As is set forth in Table 2, the dimensionless Prandtl number “Pr” of the R-134a liquid-vapor mixture having an average refrigerant quality “χ” equal to 0.70 is 1.7126. Corresponding to this value of the dimensionless Prandtl number “Pr”, we obtain from the graph of
Table 2. Data for the Calculation of the Optimal Hydraulic Diameter “do” of a Baseline Circular Passage Utilizing R-134a Refrigerant at 50° F.
The dynamic viscosity “μ” of R-134a refrigerant corresponding to an average refrigerant quality “χ” equal to 0.70 at 50° F. is also required for the calculation of “do” with the use of Equation (1). Referring to Table 2, this value is determined to be 0.0762 lbm/ft·hr.
Finally, the mass flow rate “{dot over (m)}” through the flow passage needs to be prescribed in order to compute “do” using Equation (1). Assuming that the total mass flow rate of R-134a through the evaporator is 420 lbm/hr based on the system sizing considerations and that the average number of flow passages within the evaporator tubes defining each flow pass is 300, we can determine the mass flow rate {dot over (m)} through each flow passage as 420/300=1.4 lbm/hr.
Thus given that “Φ”=0.00018, “{dot over (m)}”=1.4 lbm/hr and “μ”=0.0762 lbm/ft·hr, we find that all the information for the computation of “do” using Equation (1) is now at hand. Using these values in Equation (1) set forth above, the optimal hydraulic diameter “do” of the baseline circular flow passage is found to be equal to 0.0033 ft=0.040 in. (1 mm).
Once the optimal hydraulic diameter “do” of the baseline circular passage has been determined, the optimal hydraulic diameter “d” of any given noncircular passage can be determined. Specifically, the optimal hydraulic diameters “d” of the respective noncircular passages represented by the cross-sectional areas shown in
The example described in the following paragraphs illustrates the manner in which the optimal hydraulic diameter “d” of a noncircular passage, such as a cusped passage shown in
Referring to
Assume that the operating conditions of an evaporator utilizing tubes incorporating the cusped flow passages are identical to those of the evaporator described above in Table 2 and in paragraphs following Table 2. Thus, under these conditions the optimal hydraulic diameter “do” of the baseline circular passage can be taken as 0.040 in (1 mm) as computed above with the use of Equation (1). Given this value of “do” and the values of the ratio “d/do” for the respective cusped passages in the graph of
Another example presented below illustrates the manner in which the optimal hydraulic diameter “d” of a non-circular passage, such as a hypocycloidal passage shown in
Referring to
Assume that the operating conditions of an evaporator utilizing tubes incorporating the hypocycloidal flow passages are identical to those of the evaporator described above in Table 2 and in paragraphs following Table 2. Thus, under these conditions the optimal hydraulic diameter “do” of the baseline circular passage can be taken as 0.040 in (1 mm) as computed above with the use of Equation (1). Given this value of “do” and the values of the ratio “d/do” for the respective hypocycloidal passages in the graph of
Comparison of the data set forth in Tables 3 and 4 reveals that the optimal hydraulic diameter “do” of a circular passage calculated for a given refrigerant under a given set of operating conditions is always greater than the optimal hydraulic diameter “d” of any non-circular passage, such as a cusped passage or a hypocycloidal passage, under identical operating conditions. Furthermore, although the cusped and hypocycloidal passages are similar in shape, the magnitudes of the optimal hydraulic diameters “d” of the two types of passages are quite different. This underscores the need to establish the optimal hydraulic diameter for each flow passage to be utilized in a heat exchanger of the present invention.
The optimal hydraulic diameter is highly passage-specific and there is no universal value of the optimal hydraulic diameter applicable to all circular and noncircular passages. According to the teachings of the subject invention, the optimal hydraulic diameter ratios d/do were determined for a number of flow passages of interest as shown in
While the invention has been described with reference to exemplary embodiments, it will be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation or material to the teachings of the invention without departing from the essential scope thereof. Therefore, it is intended that the invention not be limited to the particular embodiments disclosed as the best mode contemplated for carrying out this invention, but that the invention will include all embodiments falling within the scope of the appended claims.
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