CN102401597A - Heat transfer tube for heat exchanger and heat exchanger using the same - Google Patents

Abstract

The present invention relates to a heat transfer tube for a heat exchanger and a heat exchanger using the same. The present invention provides a heat transfer tube for a heat exchanger capable of improving the performance of a water-refrigerant heat exchanger, and a heat exchanger using the same. The heat transfer tube for a heat exchanger of the present invention has: a helical inner surface groove 2b formed on the inner surface of the tube, a helical corrugated groove 3 formed on the outer surface of the tube, and a plurality of concave grooves 4 formed adjacent to the corrugated groove 3 Corrugated protrusions 5 formed on the inner surface of the tube corresponding to the corrugated grooves 3 , and recessed protrusions 6 formed on the inner surface of the tube corresponding to the recessed grooves 4 .

Description

Heat transfer tube for heat exchanger and heat exchanger using same

technical field

The present invention relates to a heat transfer tube for a heat exchanger and a heat exchanger using the same.

Background technique

For a long time, as a heat exchanger of a natural refrigerant heat pump water supply machine (hereinafter sometimes referred to as "heat pump water supply machine"), there is an internal and external double pipe consisting of an outer pipe for circulating water and an inner pipe for circulating refrigerant. Double tube heat exchanger made of tube structure. In such a double tube heat exchanger, since the inner tube through which the refrigerant flows through corrodes and opens holes, water and refrigerant will be mixed, so a device for detecting leakage of water or refrigerant and making the supply Leak detection section for water heater stop. The leak detection unit includes a leak detection tube having a leak detection groove, and by providing the leak detection tube, a heat exchanger actually having a three-layer tube structure is constituted.

On the other hand, a heat pump water heater is a machine that spends time boiling hot water at night, and the water flow rate is small and laminar. Therefore, in order to improve the performance as a heat exchanger, it is essential to improve the heat transfer performance of the water pipe that becomes the bottleneck.

As an example of a heat exchanger for a heat pump water heater aimed at improving heat transfer performance, there is a second heat transfer tube in which a plurality of heat transfer tubes are arranged in a helical shape inside a first heat transfer tube. Heat exchanger (for example, refer to Patent Document 1). It states that according to the heat exchanger described in this patent document 1, the pressure loss of water and the elution of scale components are small, and heat transfer can be promoted without using other members as heat transfer promoters.

In addition, as another example of a heat exchanger of a heat pump water heater, there is a heat exchanger in which a water pipe is used as a core pipe and a refrigerant pipe is wound from the outside (for example, refer to Patent Document 2). As the shape of the core tube, a smooth tube, a tube with grooves on the inner surface, or a structure in which a spiral plate is inserted inside the core tube is disclosed. It is described that according to the heat exchanger described in this patent document 2, it is effective in improving the easiness of manufacture and transportation, heat exchange performance, and cost reduction.

On the other hand, as a heat transfer tube most suitable for the low flow rate condition of a heat pump water heater, there is a corrugated tube with grooves on the inner surface previously proposed by the present applicant (for example, refer to Patent Document 3). The corrugated shape described in this patent document 3 can greatly improve the heat transfer performance compared with a smooth tube.

prior art literature

Patent Document 1 Japanese Patent Laid-Open No. 2004-360974

Patent Document 2 Japanese Unexamined Patent Publication No. 2002-228370

Patent Document 3 JP-A-2009-174833

Contents of the invention

However, in the case of the heat exchanger described in Patent Document 1, the process of twisting a plurality of heat transfer tubes into a helical shape itself is complicated, and the process of twisting hollow tubes that are prone to deformation such as crushing or breaking is not as good as twisting a solid tube. The process of electric wire is so easy, and it consumes production cost. In addition, the processing (structure) of the end portion of the heat exchanger that separates the first heat transfer tube from the plurality of second heat transfer tubes becomes complicated. Furthermore, when the above-mentioned leak detection part is provided, since it is necessary to form a plurality of second heat transfer tubes into a double tube structure, the production cost is high.

With regard to the heat exchanger described in the above-mentioned Patent Document 2, even if the core tube is made into a corrugated shape or a spiral plate is inserted into the core tube, the desired heat transfer performance cannot be obtained, and the production cost and pressure loss may increase. big. Also, when the core tube is a grooved tube, even if the heat transfer area is increased, the effect of the increased heat transfer area cannot be obtained in the laminar flow region where the flow velocity is low. Furthermore, due to restrictions on the manufacturing method of the inner surface grooved tube, it is difficult to form a large shape change that causes turbulent flow in the laminar flow region where the flow velocity is small.

The heat exchanger described in the above-mentioned Patent Document 3 can achieve high heat transfer performance due to the use of the bellows with grooves on the inner surface which is most suitable for natural refrigerant heat pump water heaters. In several regions, the improvement rate of heat transfer performance is not significantly reflected.

The problem to be solved by the invention

Therefore, an object of the present invention is to provide a heat transfer tube for a heat exchanger capable of improving the heat transfer performance of a water-refrigerant heat exchanger, and a heat exchanger using the same.

means of solving problems

(1) The present invention provides a heat transfer tube for a heat exchanger, characterized in that it has: a tube used as a water tube constituting the heat exchanger; a spiral inner surface groove formed on the inner surface of the tube; A spiral corrugated groove formed on the outer surface of the tube; a plurality of concave grooves formed adjacent to the corrugated groove; corrugated protrusions protruding from the inner surface of the tube corresponding to the corrugated groove; A recessed protrusion is protrudingly formed on the inner surface of the above-mentioned tube.

(2) Preferably, Hc is the corrugation depth of the corrugated groove, OD is the outer diameter of the tube, Hf is the fin height of the inner surface groove, and Hf is the maximum inner diameter of the tube. When it is ID, it satisfies 0.022(30.7×(Hc/OD)+1.13) ^(-0.5) ≤Hf/ID≤0.035.

(3) Preferably, the corrugation depth Hc of the corrugation groove and the outer diameter OD of the tube satisfy 0.03≦(Hc/OD).

(4) Preferably, 0.5≦(Pd/OD)≦1 is satisfied when the pitch of the depressed grooves is Pd and the outer diameter of the tube is OD.

(5) Preferably, βc>θ is satisfied when the helix angle formed by the corrugated grooves and the axis of the pipe is βc and the helix angle formed by the inner surface grooves and the axis of the pipe is θ.

(6) The present invention also provides a heat exchanger comprising the heat transfer tube described in any one of (1) to (5) above.

Invention effect

According to the present invention, it is possible to obtain a heat transfer tube for a heat exchanger that can improve the performance of the heat exchanger even in a use state such as a natural refrigerant heat pump water heater in which the flow rate of water is low, and a heat exchange system using the same. device.

Description of drawings

1 is an explanatory diagram schematically showing the structure of the heat transfer tube of the present invention, (a) is a plan view with a part cut away, (b) is an enlarged cross-sectional view taken along line B-B of (a), and (c) is ( The enlarged cross-sectional view of the part in the circle in the direction of C in a).

2 is a graph showing the relationship between the ratio (Hc/OD) of the corrugation groove depth Hc to the corrugation outer diameter OD and the heat transfer performance ratio to a smooth tube when the Reynolds number Re is 2000.

3 is a graph showing the relationship between the ratio (Hc/OD) of the corrugation groove depth Hc to the corrugation outer diameter OD and the pressure loss ratio with respect to a smooth tube when the Reynolds number Re is 2000.

FIG. 4( a ) is a graph showing the heat transfer performance of a helical grooved inner surface tube and a smooth tube, and FIG. 4( b ) is a graph showing an enlarged display of a part of the Reynolds number Re region.

Fig. 5 is a graph showing the results of pressure loss measurement of a helical inner grooved pipe and an inner grooved corrugated pipe.

Fig. 6(a) is a graph showing the measurement results of the heat transfer performance of a smooth pipe, a smooth inner surface corrugated pipe, an inner surface grooved corrugated pipe, and an inner surface grooved T-shaped corrugated pipe. Fig. 6(b) is a graph showing the Reynolds A graph showing an enlarged portion of the number Re region.

7 is a graph showing the results of pressure loss measurement of a smooth pipe, a smooth inner corrugated pipe, an inner grooved corrugated pipe, and an inner grooved T-shaped bellows.

8 is a graph showing the relationship between the Reynolds number in the turbulent flow region and the value δ ^* obtained by dividing the thickness of the viscous underlayer of a smooth pipe by the maximum inner diameter ID of the pipe.

9 is a graph showing the relationship between the multiple of the pipe friction coefficient of the smooth pipe and the value δ ^* obtained by dividing the thickness of the viscous underlayer by the maximum inner diameter ID of the inner surface groove.

Fig. 10(a) is a graph showing the relationship between the heat transfer performance and the laminar flow region (a region with a small Reynolds number Re), and Fig. 10(b) is a graph showing the relationship between the pressure loss and the laminar flow region (a region with a small Reynolds number Re). Relationship graph.

In the picture:

1-heat transfer tube, 2a-fin, 2b-inner surface groove, 3-corrugated groove, 4-depressed groove, 5-corrugated protrusion, 6-depressed protrusion, Hc-deep of corrugated groove, Hf-fin height, ID-the maximum inner diameter of the bellows, OD-the outer diameter of the bellows, Pc-the pitch of the corrugations, Pd-the pitch of the depressions, Ta-the axis of the tube, Tw-the wall thickness of the smooth end of the bellows, βc, θ-helix angle .

Detailed ways

Preferred embodiments of the present invention will be specifically described below based on the drawings.

(The overall structure of the heat transfer tube)

In FIG. 1 , reference numeral 1 indicating the whole shows the overall configuration of a typical heat transfer tube according to this embodiment. The heat transfer tube 1 is used as a water tube constituting a heat exchanger, and performs heat exchange between water flowing inside the heat transfer tube and refrigerant flowing outside the heat transfer tube. The illustrated heat transfer tube 1 can be suitably used, for example, as a heat exchanger for a water-refrigerant heat exchanger of a natural refrigerant heat pump type water heater.

The inner surface of the heat transfer tube 1 has, as shown in FIG. 1 , a spirally concave inner surface groove 2b by forming the fins 2a in a spiral shape over the entire direction of the tube axis Ta. On the outer surface of the heat transfer tube 1, there are adjacently formed on the same line a corrugated groove 3 which has an inclination angle different from that of the inner surface groove 2b and which is continuous in a spiral shape and constricted at the tip, and a plurality of independent rectangular concave grooves. The section of the pipe wall of the concave groove 4 has a curved wave shape. On the inner surface of the heat transfer tube 1 , corresponding to the corrugated groove 3 and the concave groove 4 , corrugated protrusions 5 with constricted tips and rectangular concave protrusions 6 are protrudingly formed. The positions in the spiral direction of the grooves formed by the depressed grooves 4 and the depressed protrusions 6 are formed every other pitch.

Here, the heat transfer tube with grooves on the inner surface that has not been processed is called a corrugated tube with grooves on the inner surface. For the heat transfer tube 1 shown in the figure, since the basic shape of the inner surface of the heat transfer tube is a corrugated tube with grooves on the inner surface, the concave The combination with corrugations has a T-shape when viewed from the front, so the illustrated heat transfer tube 1 is called a T-shaped corrugated tube 1 with inner surface grooves. In addition, the corrugated groove 3 and the corrugated protrusion 5 are called corrugations.

When the angle formed by the corrugated groove 3 of the inner surface grooved T-shaped corrugated pipe 1 and the pipe axis Ta is set as the corrugated helix angle βc, the value of the corrugated helix angle βc is in the range of 0°<βc<90° , but the helix angle βc is preferably a highly twisted shape of not less than 40° and less than 90°. The range of 40°≦βc≦82° is more preferable. When the inner surface groove helix angle formed by the inner surface groove 2b and the pipe axis Ta is set as θ, the value of the inner surface groove helix angle θ is 0°<θ<90°, but the helix angle θ preferably satisfies The condition of βc>θ. In this way, turbulence of the fluid can be promoted.

In this T-shaped bellows 1 with inner surface grooves, as shown in FIG. Let the maximum inner diameter of corrugation ") be ID, the depth of corrugation groove 3 (hereinafter referred to as "corrugation depth") be Hc, and the outer diameter of T-shaped bellows 1 with grooves on the inner surface (hereinafter referred to as "corrugation outer diameter") ) is OD, and when the wall thickness (hereinafter referred to as "corrugation wall thickness") of the end smooth portion of the T-shaped bellows with inner surface groove 1 is Tw, it is preferable to satisfy the following (a) to (c) conditions of. According to this structure, good heat transfer performance (hereinafter also referred to as "performance") can be obtained as the inner grooved T-shaped bellows 1 .

(a)0.022(30.7×(Hc/OD)+1.13) ^(-0.5) ≤(Hf/ID)≤0.035

(b) 0.04≤(Hc/OD), and

(c) OD=2Tw+ID

When the pitch of the illustrated recesses is defined as Pd, it is more preferable to satisfy the relationship of 0.5≦(Pd/OD)≦1. By adjoining the dimples and corrugations, the T-shape formed by the dimples and corrugations periodically appears with respect to the direction of the pipe axis, which is the flow direction of water, so that the leading edge effect can be repeatedly generated and the heat transfer performance can be greatly improved.

The corrugation thickness Tw and the corrugation pitch Pc are not particularly limited, but preferably satisfy the relationships of, for example, 0.4mm≤Tw≤1.7mm and 3mm≤Pc≤10mm. The material of the inner grooved T-shaped bellows 1 is not particularly limited, but it is preferable to use copper or a copper alloy, or aluminum or an aluminum alloy in consideration of thermal conductivity and mechanical strength.

FIG. 2 shows the relationship between the ratio (Hc/OD) of the corrugation depth Hc to the corrugation outer diameter OD and the heat transfer performance ratio with respect to a smooth tube when the Reynolds number Re is 2000. The corrugation outer diameter is 9.52 mm, the corrugation pitch Pc is 8 mm, and the number of corrugations is one. As can be seen from FIG. 2 , when (Hc/OD) is less than 0.04, the heat transfer performance drops sharply. Therefore, it is preferable to satisfy the relationship of 0.04≦(Hc/OD).

Fig. 3 shows the relationship between the ratio (Hc/OD) of the corrugation depth Hc to the corrugation outer diameter OD and the pressure loss (pipe friction coefficient ratio with respect to a smooth pipe) when the Reynolds number Re is 2000. The pipe friction coefficient is a dimensionless number λ specified by the relational expression of ΔP=λ×L/de×(ρv2)/2, which can be regarded as an index of pressure loss after offsetting the influence of flow channel area or fluid velocity, etc. . Here, ΔP represents the pressure loss of the inner grooved T-shaped bellows 1, L represents the length of the inner surface grooved T-shaped bellows 1, and de represents the diameter commensurate with the inner surface grooved T-shaped bellows 1 (4×flow channel area/wetting edge length), ρ represents the density of the fluid, and v represents the flow velocity of the fluid.

As can be seen from Fig. 3, when the ratio of the corrugation depth Hc to the corrugation outer diameter OD (Hc/OD) is less than 0.04, the tube friction coefficient ratio, that is, the pressure loss also drops sharply as in the heat transfer performance ratio shown in Fig. Promotes turbulence. In addition, it can be seen that when (Hc/OD) becomes 0.04 or more, the tube friction coefficient ratio continues to increase. Furthermore, it can be seen from Figure 3 that if (Hc/OD) exceeds 0.1 (0.1<(Hc/OD)), the tube friction coefficient ratio (pressure loss) exceeds the heat transfer performance, for example, when (Hc/OD)=1.1, The tube friction coefficient ratio is 4.5 relative to the heat transfer performance ratio of 4.3. Therefore, it is preferable to satisfy 0.04≦(Hc/OD)≦0.1, and the inner grooved T-shaped bellows 1 with low pressure loss and high performance can be obtained.

(Manufacturing method of T-shaped bellows with grooves on the inner surface)

As a method of manufacturing the inner grooves 2b of the inner grooved T-shaped bellows 1, general roll forming of inner grooved pipes can be employed. As one example thereof, for example, by continuously pressing the inner surface of the original pipe while a disc-shaped disc for forming concavo-convex not shown in the figure is inclined with respect to the pipe axis Ta, it is rotated. At the same time, it revolves inside the inner surface of the original pipe and draws out the original pipe at a predetermined speed, thereby forming the helical inner surface groove 2b. Of course, various processing patterns can be formed by changing the shape of the disc-shaped disk, the rotation speed, the drawing speed of the heat transfer tube, and the like.

As a method of manufacturing the corrugations (corrugation groove 3 and corrugation protrusion 5 ) and depressions (depression groove 4 and depression protrusion 6 ) of the T-shaped bellows 1 with inner surface grooves, general bellows roll forming can be used. As one example, for example, the inner surface of the original pipe is formed by aligning a disc-shaped disk (not shown) having a concave-convex portion corresponding to the corrugations and depressions on the inner surface of the original pipe while being inclined with respect to the pipe axis Ta. The outer surface of the original pipe is made to revolve on the outer surface of the original pipe while rotating, and the original pipe is drawn out at a specified speed, so that spiral corrugations and depressions can be formed. Various processing patterns can be formed by changing the shape of the disc-shaped disk, the rotation speed, the drawing speed of the heat transfer tube, and the like.

According to the above manufacturing method, the inner surface of the heat transfer tube is equipped with a spiral inner surface groove 2b, and the spiral continuous corrugated groove 3 and a plurality of independent depression grooves 4 are adjacently formed on the outer surface of the heat transfer tube, so that the heat transfer effect can be effectively obtained. The inner surface of the pipe has corrugated protrusions 5 and recessed protrusions 6 adjacent to the inner surface of the T-shaped corrugated pipe 1 with grooves formed in a T-shaped protrusion.

(Construction of heat exchanger)

The T-shaped bellows 1 with inner grooves configured as described above is used as a water pipe (inner pipe) constituting a heat exchanger. The heat exchanger (not shown) is configured to have an outer tube on the outer side of the T-shaped corrugated tube with inner surface grooves 1, so that the annular path of the refrigerant between the T-shaped corrugated tube with inner surface grooves 1 and the outer tube middle flow.

Hereinafter, referring to Tables 1 to 3 and FIGS. 2 to 9 , Example 1 and Comparative Examples 1 to 7 will be described in detail as more specific embodiments of the present invention. In addition, in this Example, one typical example among the above-mentioned embodiments was given, but, of course, the present invention is not limited to these Examples and Comparative Examples.

Referring to Figures 2 to 9, there are shown spiral pipes with grooves on the inner surface (comparative examples 1 to 4), smooth pipes (comparative example 5), corrugated pipes (comparative example 6), and corrugated pipes with inner surface grooves (comparative example 7) 1. The heat transfer performance measurement result of the T-shaped corrugated pipe with grooves on the inner surface (embodiment 1). The specifications of these heat transfer tubes are listed in Table 1 below.

Table 1

Comparative Example 7 is a heat transfer tube in which Comparative Example 4 was corrugated, and Example 1 is a heat transfer tube in which Comparative Example 1 was corrugated. Regardless of which heat transfer tube, its material is copper or copper alloy, and its corrugated outer diameter (OD) is made 9.52mm. Here, the heat transfer performance is defined as (Nu/Pr ^0.4 ) obtained by dividing the Nusselt number Nu by the 0.4th power of the Prandtl number Pr in order to offset the influence of the physical properties of the fluid. The pressure loss is also expressed by the pipe friction coefficient f of the dimensionless number Darcy.

In FIG. 4( a ), the heat transfer performance measurement results of Comparative Examples 1 to 4 (tubes with spiral inner surface grooves) and Comparative Example 5 (smooth tube) are shown in order. In FIG. 4( b ), the Reynolds number Re=5000 or less in FIG. 4( a ) is enlarged and shown. Comparative examples 3 and 4 are different from comparative examples 1 and 2 in that the heat transfer performance increases in the transition region (Reynolds number Re=2300-4000), but in the laminar flow region (Reynolds number Re=2300 or less) the heat transfer performance of comparative example 5 Performance is the same.

In Fig. 5, the pressure loss measurement results of Comparative Example 1 (spiral inner surface grooved pipe), Comparative Example 4 (helical inner surface grooved pipe) and Comparative Example 7 (inner surface grooved corrugated pipe) are shown in order . Comparative Example 4 drops sharply in the laminar flow region, contrary to Comparative Example 1. According to Fig. 4 and Fig. 5, it can be considered that the high-finned inner surface grooved tube with improved heat transfer performance in the transition region has a rectification effect in the laminar flow region.

In Fig. 6(a), a smooth pipe (comparative example 5), a smooth inner surface corrugated pipe (comparative example 6), an inner surface grooved bellows pipe (comparative example 7) and an inner surface grooved T-shaped bellows are shown in order (Example 1) heat transfer performance measurement results. In FIG. 6( b ), the Reynolds number Re=4000 or less (transition region to laminar flow region) in FIG. 6( a ) is enlarged and shown. In FIG. 7 , the pressure loss measurement results of Comparative Example 5, Comparative Example 6, Comparative Example 7, and Example 1 are shown in order.

As shown in Fig. 6, the corrugated pipe with inner surface grooves of Comparative Example 7 obtained by corrugating Comparative Example 4 (spiral-shaped inner surface grooved pipe) shown in Fig. 4 has a heat transfer performance in the laminar flow region. On the contrary, it is lower than that of the inner surface smooth bellows of Comparative Example 6. Similarly, the inner grooved bellows of Comparative Example 7 was also lower than the inner smooth bellows of Comparative Example 6 in terms of pressure loss. It is considered that the rectification effect produced by the grooves on the inner surface of Comparative Example 7 degrades the heat transfer performance of the bellows.

On the other hand, the T-shaped corrugated pipe with internal grooves in Example 1 does not reduce the heat transfer performance of the corrugated pipe in the laminar flow region, and has the same heat transfer performance as the helical internal grooved pipe of Comparative Example 1. The transition area to the turbulent flow area is the heat transfer performance above the bellows. In the transition region to the turbulent flow region, even compared with the corrugated pipe with inner grooves of Comparative Example 7, the T-shaped bellows with inner grooves of Example 1 maintains superiority in heat transfer performance. Compared with the smooth inner surface bellows of Comparative Example 6, as shown in the above Table 1, the T-shaped bellows with inner surface grooves of Example 1 has the smallest weight increase of 13%. When the Reynolds number Re is 7000, the implementation The heat transfer performance of the T-shaped bellows with grooves on the inner surface of Example 1 increased by more than 30%.

Table 2 below shows the value (Hf/ID) obtained by dividing the fin height Hf of the inner surface groove by the maximum inner diameter ID of the tube. According to Fig. 4 and Table 2, the corrugated pipe with grooved inner surface is rectified in the laminar flow region. In order not to degrade the performance of the corrugated pipe with smooth inner surface, the (Hf/ID) of the helical grooved pipe with inner surface of Comparative Example 2 needs to be Below 0.038.

Table 2

No. Hf/ID Comparative example 1 0.027 Comparative example 2 0.038 Comparative example 3 0.045 Comparative example 4 0.082

However, since the helical inner surface grooved tube of Comparative Example 1 before the corrugation of Example 1 does not improve the heat transfer performance in the transition region compared with the smooth tube, the fins are not high enough to rectify the flow in the laminar flow region. degree. However, even in the turbulent flow region (the Reynolds number Re is 4000 or more), the heat transfer performance does not improve unless the Reynolds number is high. This is due to the low fin height Hf of the grooves on the inner surface which are hidden in the turbulent boundary layer.

The turbulent boundary layer consists of a viscous or laminar substratum flowing laminarly in the immediate vicinity of the pipe wall, and a layer intermediate between laminar and turbulent flow. In order to compare the fin height of the inner surface grooves and the thickness of the viscous base layer, etc., the dimensionless number y ⁺ of the distance y from the tube wall to the tube center direction is defined as the following Mathematical Expression 1.

Mathematical formula 1

${\mathrm{<span\; class="notranslate">\; the\; y</span>}}^{<span\; class="notranslate">\; +</span>}<span\; class="notranslate">\; =</span>\frac{\mathrm{<span\; class="notranslate">\; \&rho;</span>}<span\; class="notranslate">\; \&Center\; Dot;</span>{\mathrm{<span\; class="notranslate">\; u</span>}}^{<span\; class="notranslate">\; *</span>}<span\; class="notranslate">\; \&CenterDot;</span>\mathrm{<span\; class="notranslate">\; the\; y</span>}}{\mathrm{<span\; class="notranslate">\; \&mu;</span>}}<span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 1</span>}<span\; class="notranslate">\; )</span>$

Here, ρ and μ represent the density (kg/m ³ ) and viscosity (pas) of the fluid flowing in the tube, and u ^* represents the friction velocity, and are determined by the following formula 2.

Mathematical formula 2

${\mathrm{<span\; class="notranslate">\; u</span>}}^{<span\; class="notranslate">\; *</span>}<span\; class="notranslate">\; =</span>\frac{{\mathrm{<span\; class="notranslate">\; \&tau;</span>}}_{\mathrm{<span\; class="notranslate">\; w</span>}}}{\mathrm{<span\; class="notranslate">\; \&rho;</span>}}<span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 2</span>}<span\; class="notranslate">\; )</span>$

$\frac{\mathrm{<span\; class="notranslate">\; \&pi;</span>}}{\mathrm{<span\; class="notranslate">\; 4</span>}}{\mathrm{<span\; class="notranslate">\; <figure-callout\; id="2"\; label="ID"\; filenames="HSA00000432019300021.png,HSA00000432019300061.png"\; state="\{\{state\}\}">ID</figure-callout></span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; P</span>}\mathrm{<span\; class="notranslate">\; 1</span>}<span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; P</span>}\mathrm{<span\; class="notranslate">\; 2</span>}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; =</span>\mathrm{<span\; class="notranslate">\; \&pi;</span>}<span\; class="notranslate">\; \&Center\; Dot;</span>\mathrm{<span\; class="notranslate">\; ID</span>}<span\; class="notranslate">\; \&Center\; Dot;</span>\mathrm{<span\; class="notranslate">\; L</span>}<span\; class="notranslate">\; \&CenterDot;</span>{\mathrm{<span\; class="notranslate">\; \&tau;</span>}}_{\mathrm{<span\; class="notranslate">\; w</span>}}$

Here, τ _w represents the frictional stress (Pa) of the tube wall.

The sticky bottom layer is generally in the range of 0≤y ⁺ ≤5.

Table 3 below shows the results of calculating y ⁺ corresponding to the fin height of the inner surface grooves in Comparative Examples 1 to 4 when the Reynolds number Re, which is the boundary between the turbulent flow region and the transition region, is 4000. In addition, (Hf/ID) shown in Table 2 is also described in Table 3 together.

table 3

No. Y ⁺ Hf/ID Comparative Example 1 7.8 0.027 Comparative example 2 10.6 0.038 Comparative example 3 12.7 0.045 Comparative example 4 22.9 0.092

It can be seen from Figure 4 that, according to the above Table 2, in Comparative Example 2, the Reynolds number Re of 4000 or more improves the heat transfer performance, and at this time y ⁺ is more than twice that of the viscous bottom layer.

Figure 8 shows the Reynolds number Re in the region of turbulent flow as a function of the thickness of the viscous underlayer of a smooth pipe (y at y ⁺ = 5) divided by the maximum internal diameter ID of the pipe δ ^* . According to FIG. 8 , in Comparative Example 1, the thickness of the viscous underlayer with a Reynolds number Re of 6000-7000 reaches 0.012. According to the above Table 2, the (Hf/ID) of Comparative Example 1 is 0.027, and similarly to Comparative Example 2, the performance is improved when the thickness of the adhesive underlayer is twice or more than 0.011.

According to the above formula 1, it can be seen that the thickness of the viscous bottom layer (y when y ⁺ = 5) is inversely proportional to the friction velocity u ^* . The frictional velocity u ^* is proportional to the frictional stress _τw on the pipe wall. In addition, the relationship between the frictional stress τ _w and the pressure loss ΔP (=P1-P2) of the length (section) L of the tube is expressed in Mathematical Expression 3 below.

Mathematical formula 3

∴ ${\mathrm{\&tau;}}_w=\frac{\mathrm{ID}}{4L}\mathrm{\&Delta;P}\&CenterDot;\&CenterDot;\&CenterDot;\left(3\right)$

As Darcy-Weisbach's formula (mathematical formula 4),

Mathematical formula 4

$\mathrm{<span\; class="notranslate">\; \&Delta;\&rho;</span>}<span\; class="notranslate">\; =</span>\mathrm{<span\; class="notranslate">\; f</span>}\frac{\mathrm{<span\; class="notranslate">\; L</span>}}{\mathrm{<span\; class="notranslate">\; ID</span>}}\frac{\mathrm{<span\; class="notranslate">\; \&rho;</span>}{\mathrm{<span\; class="notranslate">\; <figure-callout\; id="2"\; label="v"\; filenames="HSA00000432019300021.png,HSA00000432019300061.png"\; state="\{\{state\}\}">v</figure-callout></span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}}{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 4</span>}<span\; class="notranslate">\; )</span>$

Here, v represents the average flow velocity (m/s) of the fluid.

Substituting the above formula (4) into the above formula (3) results in the following formula (5). If the fluid temperature and flow rate are the same, the friction stress τ _w is proportional to the pipe friction coefficient.

Mathematical formula 5

${\mathrm{<span\; class="notranslate">\; \&tau;</span>}}_{\mathrm{<span\; class="notranslate">\; w</span>}}<span\; class="notranslate">\; =</span>\frac{\mathrm{<span\; class="notranslate">\; <figure-callout\; id="4"\; label="f"\; filenames="HSA00000432019300021.png,HSA00000432019300101.png"\; state="\{\{state\}\}">f</figure-callout></span>}}{\mathrm{<span\; class="notranslate">\; 4</span>}}\frac{\mathrm{<span\; class="notranslate">\; \&rho;</span>}{\mathrm{<span\; class="notranslate">\; <figure-callout\; id="2"\; label="v"\; filenames="HSA00000432019300021.png,HSA00000432019300061.png"\; state="\{\{state\}\}">v</figure-callout></span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}}{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 5</span>}<span\; class="notranslate">\; )</span>$

Fig. 9 shows the value δ ^* obtained by dividing the pipe friction coefficient f of the smooth pipe as 1, its multiple k and the thickness of the viscous bottom layer (y when y ⁺ =5) divided by the maximum inner diameter ID of the inner surface groove 2b Relationship.

The Reynolds number Re of the heat pump water heater ranges from about 1500 to 7000. Since it reaches 4000 to 7000 in the turbulent flow region, if the Reynolds number Re is 4000 and 7000, if it is conventionalized, then

(Reynolds number Re=4000)

δ ^* ＝0.018k ^-0.5

∴ 2δ ^* ＝0.036k ^-0.5

(Reynolds number Re=7000)

δ ^* ＝0.011k ^-0.5

∴ 2δ ^* ＝0.022k ^-0.5

Here, the pressure loss ratio of the bellows to the smooth pipe ( FIG. 3 ) increases linearly as 0.04<(Hc/OD), so the relationship between the pressure loss ratio k and (Hc/OD) becomes the following Mathematical Expression 6.

Mathematical formula 6

$\mathrm{<span\; class="notranslate">\; k</span>}<span\; class="notranslate">\; =</span>\mathrm{<span\; class="notranslate">\; 30.7</span>}\frac{\mathrm{<span\; class="notranslate">\; Hc</span>}}{\mathrm{<span\; class="notranslate">\; OD</span>}}<span\; class="notranslate">\; +</span>\mathrm{<span\; class="notranslate">\; 1.13</span>}<span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 6</span>}<span\; class="notranslate">\; )</span>$

Since the performance improves when Hf/ID is twice or more than the viscous base layer δ ^* , the lower limit value of (Hf/ID) at which the performance can be improved in the inner grooved corrugated pipe is the following formula 7.

math formula 7

$\mathrm{<span\; class="notranslate">\; 0.022</span>}{<span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 30.7</span>}<span\; class="notranslate">\; (</span>\frac{\mathrm{<span\; class="notranslate">\; Hc</span>}}{\mathrm{<span\; class="notranslate">\; OD</span>}}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; +</span>\mathrm{<span\; class="notranslate">\; 1.13</span>}<span\; class="notranslate">\; )</span>}^{<span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; 0.5</span>}}<span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 7</span>}<span\; class="notranslate">\; )</span>$

Preferable is Mathematical Formula 8 below.

Mathematical formula 8

$\mathrm{<span\; class="notranslate">\; 0.036</span>}{<span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 30.7</span>}<span\; class="notranslate">\; (</span>\frac{\mathrm{<span\; class="notranslate">\; Hc</span>}}{\mathrm{<span\; class="notranslate">\; OD</span>}}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; +</span>\mathrm{<span\; class="notranslate">\; 1.13</span>}<span\; class="notranslate">\; )</span>}^{<span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; 0.5</span>}}<span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&CenterDot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 8</span>}<span\; class="notranslate">\; )</span>$

In summary, in the case of the following mathematical formula 9, like the pipe with grooves on the inner surface of Comparative Example 1, even if the performance cannot be improved in the range of the Reynolds number Re of the heat pump water heater, it can be made internally The corrugated pipe with grooves on the surface uses the stirring effect of the corrugation to make the viscous bottom layer thinner, so that the performance can be improved within the range of Reynolds number Re of the heat pump water heater.

math formula 9

$\mathrm{<span\; class="notranslate">\; 0.022</span>}{<span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 30.7</span>}<span\; class="notranslate">\; (</span>\frac{\mathrm{<span\; class="notranslate">\; Hc</span>}}{\mathrm{<span\; class="notranslate">\; OD</span>}}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; +</span>\mathrm{<span\; class="notranslate">\; 1.13</span>}<span\; class="notranslate">\; )</span>}^{<span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; 0.5</span>}}<span\; class="notranslate">\; \&le;</span>\mathrm{<span\; class="notranslate">\; HF</span>}<span\; class="notranslate">\; /</span>\mathrm{<span\; class="notranslate">\; ID</span>}<span\; class="notranslate">\; \&le;</span>\mathrm{<span\; class="notranslate">\; 0.038</span>}<span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; \&Center\; Dot;</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 9</span>}<span\; class="notranslate">\; )</span>$

other embodiments

Hereinafter, referring to Table 4 and FIG. 10 , Examples 2 to 4 and Comparative Example 8 will be given and described in detail as more specific embodiments of the present invention.

By adjoining the corrugated grooves 3 of the inner surface grooved T-shaped bellows 1, the recess processing is further increased, and the performance is improved due to the leading edge effect. The specifications of the heat transfer tubes used in the test are listed in Table 4 below. Regardless of which heat transfer tube, its material is copper or copper alloy, and its corrugated outer diameter (OD) is made 10.5mm. Water having a different Reynolds number was flowed to these heat transfer tubes, and the heat transfer performance and pressure loss were measured and calculated. Here, the heat transfer performance is defined as (Nu/Pr ^0.4 ) obtained by dividing the Nusselt number Nu by the 0.4th power of the Prandtl number Pr in order to offset the influence of the physical properties of the fluid. In addition, the comparison was made using the Reynolds number Re corresponding to the water flow rate actually used in the heat pump water heater.

FIG. 10 shows a comparison of the heat transfer performance and pressure loss of Examples 2 to 4 with that of Comparative Example 8 in the laminar flow region (a region where the Reynolds number Re is small). As can be seen from FIG. 10 , if the test range is 0.5≤(Pd/OD)≤1, compared with Comparative Example 8, Examples 2 to 4 all exceed the heat transfer performance. In addition, as far as Example 3 is concerned, its performance increase rate exceeds the loss increase rate to the greatest extent, which can be said to be an optimal design.

Table 4

According to the above description, it can be seen that the T-shaped corrugated heat transfer tube 1 with grooves on the inner surface can improve the water-refrigerant heat exchanger, which is the largest bottleneck among the elements of the natural refrigerant heat pump water supply machine, when the Reynolds number Re=5000. The heat transfer performance of the water pipe can be improved, and the overall efficiency of the water supply system can be improved. In addition, due to the low pressure loss, a reduction in pump input power can be achieved. Of course, for example, the respective external shapes of the inner surface groove 2 b, the corrugated groove 3 , the concave groove 4 , the corrugated protrusion 5 , and the concave protrusion 6 are not limited to the illustrated examples.

Claims (6)

1. a heat exchanger is used heat-transfer pipe, it is characterized in that having:

The pipe that uses as the water pipe that constitutes heat exchanger;

The spiral helicine inner surface groove that forms at the inner surface of aforementioned tube;

The spiral helicine wave groove that forms at the outer surface of aforementioned tube;

The a plurality of recessed grooves that are adjacent to form with above-mentioned wave groove;

Corresponding with above-mentioned wave groove and in the outstanding corrugation lobes that forms of the inner surface of aforementioned tube;

And it is corresponding with above-mentioned recessed groove and in the outstanding depression projection that forms of the inner surface of aforementioned tube.

2. heat exchanger according to claim 1 is used heat-transfer pipe, it is characterized in that,

The ripple degree of depth of above-mentioned wave groove is made as Hc, the external diameter of aforementioned tube is made as OD, the fin height of above-mentioned inner surface groove is made as Hf, when the maximum inner diameter of aforementioned tube is made as ID, then satisfies

0.022(30.7×(Hc/OD)+1.13) ^(-0.5)≤(Hf/ID)≤0.035。

3. heat exchanger according to claim 2 is used heat-transfer pipe, it is characterized in that,

The ripple depth H c of above-mentioned wave groove and the external diameter OD of aforementioned tube satisfy 0.03≤(Hc/OD).

4. heat exchanger according to claim 1 is used heat-transfer pipe, it is characterized in that,

The spacing of above-mentioned recessed groove is made as Pd, when the external diameter of aforementioned tube is made as OD, then satisfies 0.5≤(Pd/OD)≤1.

5. heat exchanger according to claim 1 and 2 is used heat-transfer pipe, it is characterized in that,

The above-mentioned wave groove and the helical angle that axis became of aforementioned tube are made as β c, when above-mentioned inner surface groove is made as θ with the helical angle that axis became of aforementioned tube, then satisfy β c＞θ.

6. a heat exchanger is characterized in that, possesses each described heat-transfer pipe among the claim 1-5.

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