JPS62175588A - Condenser with flow path having small fluid diameter - Google Patents

Abstract

The condenser comprises a pair of flow headers (10,12), one of which has a vapour inlet whilst the other has a condensate outlet (26). Flattened distribution tubes (20) between the headers define discrete hydraulically parallel fluid pathways. Each fluid pathway has an hydraulic diameter between 0.015 to 0.040 inches. There are several condenser tubes each extending between and in fluid communication with the headers.

Description

DETAILED DESCRIPTION OF THE INVENTION Field of the Invention The present invention relates to condensing devices, and more particularly to condensers such as those used in air conditioning or refrigeration systems for condensing refrigerants.

BACKGROUND OF THE INVENTION Many condensers currently used in air conditioning or refrigeration systems use one or more corrugated conduits, or condenser tubes, on the vapor side.

The flow passages within such tubes are relatively large to avoid high resistance to steam flow and/or condensate so that excessive pressure differentials do not exist from the steam inlet to the outlet, which would necessarily increase the energy requirements of the system. It is.

This ultimately means that the air side of the tube is relatively large. Because the air side of the tube is relatively large, a relatively large portion of the front surface area on the air side is blocked by the tube.
Thereby, there is less area available for arranging air side fins to enhance heat transfer.

Eventually, the pressure drop on the air side to maintain the desired rate of heat transfer becomes undesirably large, and the system energy requirements in passing the required volume of air through the air side of the condenser are proportionately large. becomes undesirably large. The present invention is intended to solve the above problem.

SUMMARY OF THE INVENTION A primary object of the present invention is to provide a new and improved condenser for use in air conditioning or refrigeration systems. In particular, it is an object of the invention to have a small frontal area on the air side blocked by the condenser tubes, without increasing the air side pressure drop and without increasing the steam and/or condensate side pressure drop; It is an object of the present invention to provide a condenser that allows the air side heat transfer surface to be increased.

The present invention achieves this object in an embodiment of a condenser consisting of a pair of spaced headers, one having a steam inlet and the other having a condensate outlet.

In the condenser, condenser tubes extend between and are in communication with the pair of headers. The condenser tubes define a plurality of substantially independent parallel fluid flow paths between the headers. The fluid diameter of each channel is approximately α015 to 1040 inches (approximately C
L4 to 1.0 melimeters).

In a preferred embodiment, a sufficient number of such condenser tubes extend between the headers in mutually parallel flow conditions to avoid high resistance to condensate and/or vapor flow.

The present invention contemplates the use of flat tubes for the condenser tube.

In a particularly preferred embodiment, a plurality of flow passages are defined within each condenser tube by corrugated spacers housed within the condenser tube.

Fins are provided on the outside of the condenser tubes and extend between adjacent condenser tubes.

The header of the present invention is defined by a generally cylindrical tube having opposing openings, such as slots, for receiving each end of the condenser tube.

DESCRIPTION OF THE PREFERRED EMBODIMENTS A condenser according to the invention is illustrated in FIG. 1 and includes generally parallel headers 1o and 12 spaced on either side. In accordance with the present invention, headers 10 and 12 are preferably made from generally cylindrical tubing. Opposing sides of the tubes are provided with a generally parallel row of slots or openings 14 for receiving corresponding ends 16 and 18 of condenser tubes 20.

Preferably, in the portion indicated by number 22 between the slots of each of the headers 10 and 12.

A slightly spherical dome is provided for improved pressure resistance, as more fully described in U.S. Patent Application No. 722.653. One end of the header 10 is closed by a lid 24 that is brazed or fused thereto. At the opposite end there is a part 26 to which a conduit 28 can be connected.
soldered or welded.

The lower end of the header 12 is opened by a soldered or welded lid 30 similar to the lid 24, while a component 32 is welded or soldered in place to its upper end. Depending on the orientation of the condenser, one of the parts 26 and 32 acts as a vapor inlet and the other as a condensate outlet. For the orientation of FIG. 1, part 26 would act as a condensate outlet.

A plurality of condenser tubes 20 are connected to the headers 10 and 1 in mutual communication.
Distract between the two. The tubes 20 are parallel to each other both geometrically and in the flow direction. Corrugated fins 34 are disposed between adjacent tubes 20, although flat fins could be used if desired. Upper and lower channel members 36 and 58 extend between headers 10 and 12 and are connected to header 1 by appropriate means to provide rigidity to the system.
0 and 12.

As shown in FIG. 1, the vessels 20 are flat tubes with corrugated spacers or inserts 40 included therein.

The spacer 40 has a cross-section as shown in FIG. 2 and, as can be seen, the alternating crests contact the inner wall 42 of the tube 20 along its entire length and are joined thereto by fillets 44. . As a result, substantially independent, parallel flow channels 4
6.48.50.52.54.56.58, and 60 are provided inside each tube 20. That is, there is substantially no flow from one such flow path to the adjacent flow path on each side.

This results in adjacent channels 46.48.5o.

This means that the six walls dividing 52, 54, 56, 58 and 60 are connected to both sides of the flat tube 20 over their entire length. That is, there are no gaps filled with fluid with lower thermal conductivity.

As a result, heat transfer from the fluid to the outside of the tube via the walls dividing the plurality of fluid flow paths mentioned above is maximized. Additionally, it is believed that in independent channels of the dimensions described, the desired effects of heat transfer based on surface tension phenomena are exploited.

A second advantage resides in the fact that condensers such as the present invention are used on the outlet side of the compressor and are therefore subjected to extremely high pressures. Conventionally, these high pressures are applied to the inside of tube 20. In that case, so-called I+play II fins are used instead of the corrugated fins 34 shown. The plate fins tend to constrain the tube 20 and thereby support it against internal pressures used in condenser applications. Conversely, corrugated fins, such as those shown at 54, are incapable of supporting tube 20 against substantial internal pressure. However, according to the invention, the desired support effect in the corrugated fin heat exchanger is achieved by the fact that the insert 40 and its corrugations are joined along the entire length of the inner wall 42 of the vessel 20. . This coupling causes various regions of the insert 40 to be placed in tension when the tube 20 is pressurized to absorb forces that tend to expand the tube 20 caused by pressure inside the tube 20.

One means by which g20 including insert 40 may be formed is described in US Pat. No. 740.000, and a particularly preferred such means is described in US Pat. No. 88,223.

According to the invention, each channel 48.50.52.54.56
and 58 and, depending on the shape of the insert 40, the flow path 4.
6 and even 60 similarly have fluid diameters ranging from about 1015 to α040 inches (about CL4-1.02 millimeters). According to current common assembly techniques known in the art, a fluid diameter of about 0.035 inches (about cL9 mm) optimizes maximum heat transfer efficiency and ease of assembly. fluid diameter
diameter ) is as conventionally defined, ie, the cross-sectional area of each channel multiplied by 4 and divided by the wetted perimeter of the corresponding channel.

The hydraulic diameter values given are for the R-12 system condenser. Somewhat different values may be expected in systems using different refrigerants.

Within the above size range, it is desirable to make the tube dimension transverse to the direction of air flow through the core as small as possible. This ultimately provides a larger front area on which fins, such as fins 34, can be placed in the core without unfavorably increasing the air side pressure drop in order to obtain better heat transfer coefficients. In some instances, minimizing bone width may allow for the placement of one or more additional tube rows.

In this regard, the preferred embodiment contemplates the use of a tube provided with a separate spacer, as illustrated in FIG. 2, as opposed to an extruded tube having a passageway of a given fluid diameter.

making mass production of condensers economically viable at this time;
With current extrusion technology, the tube wall thickness is generally greater than that required to support a given pressure using tubes and spacers such as those described herein. Ultimately, the total width of such an extruded tube for a given fluid diameter is
It is somewhat larger than if a tube and spacer combination were used, which is undesirable for the reasons mentioned above. Nevertheless, the present invention also contemplates the use of extruded tubes with fluid diameter channels within the aforementioned size ranges.

It is also desirable to make the ratio of the tube outer circumference to the inner tube collapse circumference as small as possible without making the flow path sufficiently small that refrigerant cannot readily flow therethrough. This reduces the resistance to heat transfer on the steam and/or conduit side.

The many benefits of the present invention are apparent from the data illustrated in FIGS. 3-6 and the discussion below. for example,
Figure 3 shows the condenser core product as the prior art.
The heat transfer coefficient versus cavity or fluid diameter in inch dimensions for air flows varying from 450 to 3200 standard cubic feet per minute is plotted on the right. On the left side of this data is shown a computer output curve based on a heat transfer model created using empirically obtained data for a core made in accordance with the present invention. ′
The curve designated A' is approximately 24 inches long (approximately 61 centimeters), tube wall thickness 1015 inches (approximately α4 mm), and tube major dimensions [L 532 inches (approximately 1 & 5 mm); square inch (approximately 5.08
Figure 1 represents the heat transfer in the air flow described above for a core as shown in Figure 1 having a frontal area of 7 inches squared. The inlet air temperature here is 110"tl approximately 43.3℃)
, the inlet temperature is 180″F (approximately 82.2°C), and R
The pressure on the -12 system is 255 psig% and the subcooling temperature of the effluent refrigerant after condensation is 2'F (approximately -16
7℃). Between the core tube and the fin,
There are 18 fins per inch (approximately 25 centimeters). The dimensions of the fin are cL625 inches
α540 inches x 1006 inches (approximately 15.9 mm x 117 mm x 0.15 mm).

The curve designated If B If shows the same relationship for the same core, except that the channel length in the vessel tube is doubled, ie the tube machine is halved and the tube length is doubled. As seen in FIG. 3, heat transfer through use of the present invention ranges from about 0.015 inches to about 0.040 inches (about
For a fluid diameter of 1 flI from 1 mm to about t02 milJ meters), it advantageously and substantially increases, with some variation depending on the airflow.

In FIG. 4, actual test data for a core according to the invention having the dimensions listed in Table 1 below are shown.
A comparison is made with actual test data for a conventional condenser core similar to the present invention. Data for the conventional core is also listed in Table-1.

Cores made according to the present invention and conventional cores are
1800 standard cubic feet per minute (as shown in Figure 4)
Both have the same design point of a heat transfer rate of 26,0 OOBTU per hour (approximately 540 standard cubic meters per minute). However, the actual observed equivalence of two cores is 28.0 OOBTU and 2.0 OOBTU/min. , ooo standard cubic feet (approximately 600 standard cubic meters per minute). These parameters can be used for comparison purposes.

Referring to the l D I+ and N E 11 curves representing a conventional condenser and the present invention, respectively, it will be appreciated that the refrigerant flow rate for both is approximately the same over a wide range of air flow values. For this test and the pond test illustrated in Figures 4 through 6, 180°F (approximately 82.2°C)
A 1-12 system was applied to the condenser population at , 235 psiH. The outflow refrigerant is 2″F (approximately -16,7″C)
) was supercooled. The inlet air temperature to the condenser is 11
The temperature was below 0 (approximately 416°C).

The fact that the refrigerant side pressure drop across the conventional core is greater than that across the core according to the present invention means that the energy consumed by the compressor in the conventional system is greater than that according to the present invention. It also suggests something big.

Curves II F 11 and 11 G11 are also for the conventional condenser and the inventive condenser, respectively, and show comparable heat transfer over the same air flow range.

Tracks t, I If HII and II JII are for a conventional condenser and a condenser of the present invention, respectively, and illustrate the considerable difference in refrigerant pressure drop across the condenser. These demonstrate one benefit of the present invention. Due to the much lower pressure drop across the condenser according to the invention, the average temperature of the refrigerant, whether in condensate or vapor form, is higher than in conventional condensers. As a result, for the same inlet air temperature there will be a larger temperature difference, which enhances the rate of heat transfer according to Fourier's law.

In the core according to the invention, the air side pressure drop is also lower than in conventional cores. This is based on two factors: the core depth is smaller and the free flow area occupied by the tube is larger. This ultimately saves the fan energy required to direct the desired airflow through the core. Moreover, the curve If
The heat transfer coefficient remains substantially the same, as indicated by F If and 1IGI'.

Cores according to the invention retain less refrigerant than conventional cores. Therefore, the core of the present invention reduces system requirements for refrigerant. Similarly, since the core of the present invention has a small depth, it requires less space for installation.

It can be seen from the data shown in the table and FIG. 4 that the core according to the present invention is significantly more robust than the conventional core. Thus, in Fig. 5, the amount of heat transfer per one bond (about n4s'5kl) of the conventional core (curve II K
H) and the amount of heat transfer per bond (curve If L'') of the condenser of the present invention are compared at various air velocities.

Therefore, FIG. 5 shows that by using the condenser of the present invention,
This demonstrates that significant weight savings can be made in the system without sacrificing heat transfer capabilities.

Curve "M" shown in FIG. 6 illustrates the conventional core air side pressure drop for various air flows, tj
Jt il II N If illustrates the air side pressure drop of the core of the present invention. It will be appreciated that this reduces the air side pressure drop and therefore the fan energy when using a core according to the invention.

Although the present invention has been described above based on embodiments, it should be noted that many changes may be made within the present invention.

FIG. 1 is an exploded perspective view of a condenser according to the present invention.

FIG. 2 is an enlarged cross-sectional view of a condenser conduit that may be used with the present invention.

FIG. 3 is an expected performance graph prepared by plotting heat transfer versus 0 Jil (fluid) diameter of a condenser having the same +FJ surface, one constructed by a prior art design and the other in accordance with the present invention. .

FIG. 4 is a graph comparing the conventional product and the present invention in (a): heat transfer amount, (b): refrigerant flow rate, and (C): air passage amount with respect to refrigerant pressure drop.

FIG. 5 is a graph comparing the prior art product and the present invention based on air velocity versus heat transfer per bond of material used in the manufacture of each core.

FIG. 6 is a graph comparing the prior art product to the present invention by plotting air velocity versus pressure drop across the air side of the condenser.

The names of the main parts in the figure are as follows.

10.12: Header 20: Condenser tube 34: Corrugated fin 40: Corrugated spacer 46.48.50.52.54.56.58.60: Engraving of parallel flow channel drawing (no change in content) Quarpt and Shun 17 University) Sixth Junior High School, -5CFMF'/
G4 F'lG, 5 Procedure correction band (method) to be corrected % formula % Display of case Patent application No. 231359 of 1985 Title of invention Case with a person who corrects a condenser equipped with a flow path with a small fluid diameter Related patent applicant name
Modine Manufacturing Company 103 Address: Oil and Fat Industry Hall, 3-13-11 Nihonbashi, Chuo-ku, Tokyo Telephone number: 273-6436 Name (6781) Patent attorney Motoi Kurauchi
Hiroshi, 1 ken address Same ``Yu drawing 1 copy Details of amendments Engraving of the drawing as attached (no change in content)

Claims (1)

[Claims] 1. A pair of spaced apart headers, one having a steam inlet and the other having a condensate outlet, and a tube extending parallel to the headers and communicating with each of the headers. and,
The tube has a fluid diameter of about 0.015 between the headers.
A condenser defining a plurality of parallel flow passages ranging from 0.040 inches to 0.040 inches. 2. The condenser according to claim 1, wherein each tube defines a plurality of flow paths. 3. A condenser according to claim 1, wherein the tubes are flat tubes, and the plurality of flow paths in each flat tube are defined by corrugated spacers housed in the tubes. 4. The condenser according to claim 1, wherein the tube is a condenser tube. 5. The condenser tubes are plural, each extending between a pair of headers,
5. A condenser as claimed in claim 4, wherein each condenser tube and each fluid flow path are in parallel flow to each other. 6. The condenser according to claim 5, comprising fins on the outside of the condenser tube. 7. The condenser according to claim 5, comprising fins extending between the outer sides of adjacent condenser tubes. 8. The condenser of claim 5, wherein the header is defined by a generally cylindrical tube and has opposing openings for receiving each end of the condenser tube. 9. A pair of spaced apart headers, one having a steam inlet and the other having a condensate outlet, and having opposite ends extending in parallel flow between the headers, the ends communicating with the header. a plurality of flat tubes disposed within corresponding elongated slots, and a plurality of parallel fluid flows within each flat tube between the headers having fluid diameters ranging from about 0.015 to 0.040 inches. a corrugated insert in each flat tube defining a channel, and fins extending between the outer sides of adjacent flat tubes, each header having a row of elongated slots; A condenser comprising elongated slots in one header opposed to elongated slots in the other header.