JP4420117B2 - Heat exchanger tube for heat exchanger and heat exchanger using the same - Google Patents

Description

  The present invention relates to a heat exchanger tube for a heat exchanger and a heat exchanger using the same, and more particularly to a heat exchanger tube for a water-refrigerant heat exchanger of a natural refrigerant heat pump type water heater and a heat exchanger using the same.

  As a heat exchanger of a natural refrigerant heat pump water heater (hereinafter sometimes simply referred to as “heat pump water heater”) that has been well known, an outer pipe through which water flows and an inner pipe through which refrigerant flows are used. There is a double-tube heat exchanger consisting of double tubes.

  In the case of such a double-pipe heat exchanger, if a hole due to corrosion opens in the inner pipe through which the refrigerant flows, water and the refrigerant will be mixed, so the leakage of water or refrigerant is detected and the device is stopped. In many cases, a leak detection unit (leakage detection pipe having a leak detection groove) is provided (a triple pipe structure is provided by providing a leak detection pipe).

  On the other hand, natural refrigerant heat pump water heaters boil hot water at night, and the flow rate of water is small and the flow becomes laminar. Therefore, in order to improve the performance as a heat exchanger, it is a bottleneck. It is essential to improve the heat transfer performance of the water pipe.

  The heat exchanger for the purpose of improving the heat transfer performance is 0.04 ≦ Hc / OD, where the corrugated groove depth is Hc and the corrugated outer diameter is OD, and the angle between the corrugated groove and the tube axis Ta is There is a heat exchanger composed of a corrugated tube with βc ≧ 40 ° when the twist angle βc is set, and an outer tube disposed outside the corrugated tube (see Patent Document 1).

  According to the heat exchanger described in Patent Document 1, a heat exchanger tube for a heat exchanger that can improve the heat transfer performance of the heat exchanger even in a use form where the flow rate of water is small, such as a natural refrigerant heat pump type hot water heater, and the heat exchanger tube are used. It is described that a heat exchanger can be obtained.

  Further, there is a heat transfer tube in which a continuous groove in the tube axis direction or a spiral shape is formed on the inner surface of the tube, and the cross-sectional shape cut by a plane including the tube axis is a waveform (see Patent Document 2).

  According to the heat transfer tube described in Patent Document 2, it is described that an internally grooved tube excellent in bending workability can be manufactured continuously at a low cost.

JP 2007-218486 A Japanese Patent Laid-Open No. 61-125592

  The heat transfer tube described in Patent Document 1 can increase the performance to a level equal to or greater than that of the smooth tube with the same weight as the smooth tube. In addition, a corrugated shape that takes pressure loss into consideration can be obtained and is optimized as the outer surface shape of the corrugated tube, but in order to achieve further performance improvement without increasing the weight as much as possible, the corrugated shape It is necessary to improve the inner surface shape of the tube.

  The heat transfer tube described in Patent Document 2 is a corrugated tube in which a spiral continuous groove is provided on the inner surface of the tube. If the corrugated tube described in Patent Document 1 is manufactured, the performance of the heat transfer tube is higher than that of the corrugated tube described in Patent Document 1. there is a possibility.

  However, as a result of studies by the present inventors, it has been found that even if the inner groove is simply formed in the corrugated pipe, the performance of the corrugated pipe may be lowered depending on the inner groove shape. It was also found that there was a problem that the weight increased more than the performance.

  The objective of this invention is providing the heat exchanger tube for heat exchangers which can improve the heat transfer performance of a water-refrigerant heat exchanger, and a heat exchanger using the same.

In order to achieve the above object, the present invention is a corrugated heat transfer tube used as a water tube constituting a heat exchanger, wherein a corrugated inner surface groove is formed on the inner surface of the corrugated tube, When the corrugated groove depth of the shape is Hc and the corrugated outer diameter is OD, 0.04 ≦ Hc / OD ≦ 0.1 is satisfied, the fin height of the inner groove is Hf, and the maximum inner diameter of the tube is ID. Then
0.022 {30.7 × (Hc / OD) +1.13} ^(−0.5) ≦ Hf / ID ≦ 0.037
A heat exchanger tube for a heat exchanger characterized by satisfying

  Moreover, this invention provides the heat exchanger provided with said heat exchanger tube for heat exchangers.

  ADVANTAGE OF THE INVENTION According to this invention, the heat exchanger tube for heat exchangers which can improve the heat transfer performance of a water-refrigerant heat exchanger, and a heat exchanger using the same can be obtained.

  A preferred embodiment of the present invention will be described below in detail with reference to the accompanying drawings.

[First embodiment of the present invention]
(Configuration of heat transfer tube)
FIG. 1 is an explanatory view showing the structure of a heat exchanger tube for a heat exchanger according to a first embodiment of the present invention, FIG. 1 (a) shows an overall view, and FIG. 1 (b) shows FIG. ), And FIG. 1C is an enlarged cross-sectional view in the region of the ellipse A in FIG.

  The heat transfer tube according to the present embodiment is an internally grooved corrugated tube 10 that is corrugated with a single thread after processing a spiral continuous groove on the inner surface of a smooth tube to form an inner surface groove 2, and a heat exchanger ( For example, it is used as a water pipe constituting a water-refrigerant heat exchanger for a heat pump water heater.

  That is, heat exchange is performed between the water flowing in the inner grooved corrugated pipe 10 and the refrigerant flowing outside the inner grooved corrugated pipe 10. A corrugated tube generally refers to a tube having a corrugated spiral structure on its inner and outer surfaces.

In the corrugated pipe 10 with an inner groove according to the present embodiment, when the depth of the corrugated corrugated groove 1 is Hc and the corrugated outer diameter is OD, Hc / OD representing the ratio of irregularities to the outer diameter is 0.04 ≦ When Hc / OD is satisfied, the fin height of the inner surface groove 2 is Hf, the terminal smooth portion thickness Tw, and the maximum inner diameter of the tube is ID (= OD-2Tw),
0.022 {30.7 × (Hc / OD) +1.13} c ≦ Hf / ID ≦ 0.037
It is characterized by satisfying. (Refer to [About the upper limit of the mathematical expression of the present invention] and [About the lower limit of the mathematical expression of the present invention] for detailed examination and consideration and derivation of the mathematical expressions corresponding to them.)

  By setting Hc / OD and Hf / ID within the above ranges, good heat transfer performance can be obtained. Desirably, 0.04 ≦ Hc / OD ≦ 0.1, and more desirably 0.04 ≦ Hc / OD ≦ 0.07. Moreover, it can be set as a low pressure loss by making Hc / OD into the said range.

  Further, if the angle formed by the corrugated groove 1 of the corrugated pipe 10 with the inner surface groove and the pipe axis Ta is the twist angle βc, the range of values that the twist angle βc can take is 0 ° <βc <90 °. βc is preferably a high twist shape of 40 ° or more. More desirably, 40 ° ≦ βc ≦ 82 °.

  Thereby, the turbulent flow of the fluid over the unevenness can be promoted.

  The terminal smooth portion thickness Tw and the corrugated pitch Pc of the corrugated pipe 10 with the inner surface groove are not particularly limited, but for example, 0.4 mm ≦ Tw ≦ 1.7 mm, 3 mm ≦ Pc ≦ 15 mm can be used. .

  In addition, the material is not particularly limited, but copper, copper alloy, aluminum, aluminum alloy, or the like is preferably used in consideration of thermal conductivity and mechanical strength.

[Second Embodiment of the Present Invention]
(Configuration of heat transfer tube)
FIG. 2 is an explanatory view showing the structure of a heat exchanger tube for a heat exchanger according to a second embodiment of the present invention.

  The corrugated pipe 20 with an inner groove according to the present embodiment is a corrugated pipe in which the inner grooved corrugated pipe 10 according to the first embodiment is processed with a single line, whereas the corrugated pipe is processed with three lines. It is a pipe and is used as a water pipe constituting a heat exchanger. As the number of strips increases, the processing speed increases, so there is a great cost advantage.

In the heat transfer tube of FIG. 2 as well, as in FIG. 1, when the corrugated corrugated groove depth is Hc and the corrugated outer diameter is OD, Hc / OD representing the ratio of irregularities to the outer diameter is 0.04 ≦ Hc. / OD is satisfied, the fin height of the inner groove is Hf, the terminal smooth portion thickness Tw, and the maximum inner diameter of the tube is ID (= OD-2Tw).
0.022 {30.7 × (Hc / OD) +1.13} ^(−0.5) ≦ Hf / ID ≦ 0.037
To satisfy. (Refer to [About the upper limit of the mathematical expression of the present invention] and [About the lower limit of the mathematical expression of the present invention] for detailed examination and consideration and derivation of the mathematical expressions corresponding to them.)

  The twist angle βc tends to be smaller in the case of three-strip processing than in the single-strip processing, but the inner grooved tube is manufactured by reducing the interval between adjacent corrugated grooves 1, that is, the corrugated pitch Pc. A difficult twist angle of 40 ° or more can be realized. It was difficult to increase the twist angle of the inner groove of the inner grooved tube.

  Next, the heat exchanger provided with the said corrugated pipe is demonstrated.

[Third embodiment of the present invention]
(Configuration of heat exchanger)
FIG. 3 is an explanatory view showing the structure of a heat exchanger according to the third embodiment of the present invention.

  A heat exchanger (double-pipe heat exchanger) 100 according to the present embodiment uses the heat transfer tube (for example, corrugated tube 10 with an inner surface groove) according to the embodiment of the present invention as an inner tube, and the outside thereof. The outer tube 11 is provided, and the inner surface grooved corrugated tube 10 and the outer tube 11 are formed so that the refrigerant flows through the annular path.

[Fourth embodiment of the present invention]
(Configuration of heat exchanger)
FIG. 4 is an explanatory view showing the structure of a heat exchanger according to the fourth embodiment of the present invention.

  The heat exchanger (triple tube heat exchanger) 200 according to the present embodiment has the heat transfer tube (for example, corrugated tube 10 with an inner surface groove) according to the above-described embodiment as an inner tube, and the outer periphery thereof. A leak detection tube 12 made of a smooth tube is disposed so as to form a leak detection portion (leakage detection groove 13), and an outer tube 11 is disposed outside the leak detection tube 12, and the leak detection tube 12 and the outer It is formed so that the refrigerant flows through the annular path between the tubes 11.

[Fifth and Sixth Embodiments of the Present Invention]
(Configuration of heat exchanger)
FIG. 5 is an explanatory view showing the structure of a heat exchanger according to the fifth embodiment of the present invention. Moreover, FIG. 6 is explanatory drawing which shows the structure of the heat exchanger which concerns on the 6th Embodiment of this invention.

  The heat exchangers 300 and 400 shown in FIGS. 5 and 6 are obtained by forming the outer tube in the heat exchanger of FIGS. 3 and 4 into a corrugated shape to form the corrugated outer tube 21, thereby improving flexibility. Can be made.

[Seventh embodiment of the present invention]
(Configuration of heat exchanger)
FIG. 7 is an explanatory view showing the structure of a heat exchanger according to the seventh embodiment of the present invention.

  The heat exchanger 500 according to the present embodiment includes a spiral heat transfer tube 31 for circulating refrigerant along the corrugated groove of the heat transfer tube (for example, the corrugated tube 10 with an inner surface groove) according to the above-described embodiment of the present invention. Is wound around. If necessary, the corrugated groove and the helical heat transfer tube 31 may be fixed by brazing or the like.

  In the heat exchanger 500, heat exchange is performed between the water flowing in the inner grooved corrugated pipe 10 and the refrigerant flowing in the spiral heat transfer pipe 31 in contact with the outer periphery of the inner groove corrugated pipe 10. Further, by winding the helical heat transfer tube 31 along the corrugated groove 1, the effective contact area (effective heat transfer area) between the inner grooved corrugated tube 10 and the helical heat transfer tube 31 can be increased.

[Other Embodiments of the Present Invention]
As an embodiment of the present invention, there are various forms in addition to the above first to seventh embodiments, and examples include the following forms.

  (1) Although the first and third corrugated pipes have been described, the number may be two or four or more. 1 to 3 corrugated pipes are desirable in that it is easy to achieve a high twist angle, which is difficult with an internally grooved pipe.

[Significance of numerical limitation]
FIG. 8 shows the relationship between the Hc / OD and the heat transfer performance of the corrugated pipe without the inner groove when the Reynolds number Re is 1000 (heat transfer performance ratio with respect to the smooth pipe).

  In FIG. 8, the corrugated tube has an outer diameter of 9.52 mm, a pitch of 8 mm, and a number of strips of one.

  As is apparent from FIG. 8, when Hc / OD is less than 0.04, the heat transfer performance is drastically reduced. Therefore, it is desirable to satisfy 0.04 ≦ Hc / OD.

  FIG. 9 shows the relationship between the twist angle βc of the corrugated pipe not provided with the inner groove and the heat transfer performance (heat transfer performance ratio to the smooth pipe) when the Reynolds number Re is 1000. In FIG. 9, the Hc / OD and the number of strips of the corrugated tube are 0.1 (= Hc / OD) and 1 strip.

  As is clear from FIG. 9, if Hc / OD = 0.1, the heat transfer performance is about 1.5 times higher than the smooth tube ratio even if the twist angle βc is small (for example, βc = 35 °). , Βc ≧ 40 ° makes it possible to improve the heat transfer performance more than twice the smooth tube ratio by making the twist angle high.

  FIG. 10 shows the relationship between the Hc / OD of the corrugated pipe and the pipe friction coefficient (the pipe friction coefficient ratio with respect to the smooth pipe) when the Reynolds number Re is 1000.

Here, the pipe friction coefficient is a dimensionless number f defined by the relational expression of ΔP = f × L / de × (ρv ² ) / 2, and offsets the influence of the flow path area, fluid flow velocity, and the like. It can be regarded as an indicator of pressure loss.

  ΔP is the pressure loss of the heat transfer tube, L is the heat transfer tube length, de is the equivalent diameter of the heat transfer tube (4 × channel area / wetting edge length), ρ is the fluid density, and v is the fluid flow velocity.

  As can be seen from FIG. 10, when Hc / OD is less than 0.04, the ratio of pipe friction coefficient decreases rapidly as well as the ratio of heat transfer performance, and turbulence cannot be promoted.

  On the other hand, when the Hc / OD is 0.04 or more, the pipe friction coefficient ratio (that is, pressure loss) continues to increase. Furthermore, when Hc / OD exceeds 0.1 (0.1 <(Hc / OD)), it can be seen that the pipe friction coefficient ratio exceeds the heat transfer performance ratio (see FIG. 8) (for example, Hc / OD). At OD = 0.11, the heat transfer performance ratio is 4.3 in FIG. 8 and the pipe friction coefficient ratio is 4.5 in FIG.

  Therefore, it is desirable to satisfy 0.04 ≦ Hc / OD ≦ 0.1, and a high-performance corrugated tube with low pressure loss can be provided.

  FIGS. 11 to 14 show a spiral inner surface grooved tube (Reference Examples 1 to 3), a smooth tube (Comparative Example 1), an inner surface smooth corrugated tube (Comparative Example 2), and an inner surface grooved corrugated tube (Comparative Example 3, Implementation). It is a heat-transfer performance measurement result of Example 1).

  The specifications are shown in Table 1.

  In Comparative Example 3, corrugation is applied to Reference Example 3, and in Example 1, corrugation is applied to Reference Example 1.

  All the heat transfer tubes were made of phosphorous deoxidized copper and the outer diameter (OD) was 9.52 mm.

Here, the heat transfer performance is defined as the Nusselt number Nu divided by the 0.4th power of the Prandtl number Pr in order to offset the influence of the physical properties of the fluid (Nu / Pr ^0.4 , in the following examples) The same). Similarly, the pressure loss is expressed by a Darcy tube friction coefficient f which is a dimensionless number.

  FIG. 11 shows the heat transfer performance measurement results of the spiral inner grooved tubes of Reference Examples 1 to 3 and the smooth tube of Comparative Example 1. FIG. 11B is an enlarged view of the region of Reynolds number Re5000 or less in FIG.

  In FIG. 11, unlike Reference Examples 1 and 2, the performance of Reference Example 3 is improved in the transition region (Reynolds number Re: 2300 to 4000), but in the laminar flow region (Reynolds number Re: 2300 or less), the smooth tube ( It is equivalent to Comparative Example 1).

  FIG. 12 shows pressure loss measurement results of Reference Examples 1 to 3 and Comparative Example 1.

  Reference Example 3 rapidly decreases in the laminar basin and is reverse to Reference Example 1.

  From FIG. 11 and FIG. 12, it can be said that the inner grooved pipe having a high fin whose performance is improved in the transition region has a rectifying action in the laminar flow region. On the other hand, it can be said that the inner grooved pipes with low fins as in Reference Example 1 and Reference Example 2 do not have a rectifying action in a laminar flow region, like a smooth pipe.

  FIG. 13 shows the heat transfer performance measurement results of the smooth tube (Comparative Example 1), the inner surface smooth corrugated tube (Comparative Example 2), and the inner surface grooved corrugated tube (Comparative Example 3, Example 1).

  FIG.13 (b) is an enlarged view of Reynolds number Re4000 or less (transition zone-laminar flow zone) of Fig.13 (a).

  FIG. 14 shows the pressure loss measurement results.

[Upper limit of numerical formula of the present invention]
As shown in FIG. 13 (b), the corrugated pipe with inner groove of Comparative Example 3 (corrugated to Reference Example 3 of FIG. 11) has a performance in the laminar flow area rather than the inner smooth corrugated pipe of Comparative Example 2. It is falling. Similarly, as shown in FIG. 14, the pressure loss is also lower in the laminar flow region in the inner grooved corrugated pipe of Comparative Example 3 than in the inner smooth corrugated pipe of Comparative Example 2.

  This can be said that the heat transfer performance that should be improved by the turbulent flow effect due to the corrugated shape is offset by the rectifying action by the high inner groove of the fin of Comparative Example 3, and the performance of the corrugated pipe is reduced.

  On the other hand, the corrugated pipe with the inner surface groove of Example 1 has the same performance as the inner surface smooth corrugated pipe of Comparative Example 2 without degrading the performance of the corrugated pipe in the laminar flow region, and Comparative Example 2 in the transition region to the turbulent region. The performance is better than the inner smooth corrugated tube. Compared with Comparative Example 3 in this region, Example 1 shows superior performance.

  Further, Example 1 has the smallest weight increase of 13% (Table 1) as compared with the inner surface smooth corrugated pipe of Comparative Example 2, and the performance is improved by 30% or more at the Reynolds number of 7000.

  Table 2 shows a value (Hf / ID) obtained by dividing the fin height Hf by the maximum inner diameter ID of the tube.

  From the above description and Table 2, in order to prevent the corrugated pipe with the inner groove from being rectified in the laminar flow region and not to lower the performance improvement due to the corrugated shape, the Hf / ID of Reference Example 2 is 0.037 or less. There must be.

[About the lower limit of the mathematical formula of the present invention]
By the way, since the reference example 1 before corrugating of Example 1 does not improve the performance of the smooth tube in the transition region, the fins are not so high as to rectify the flow in the laminar flow region. However, even in a turbulent region (Reynolds number Re: 4000 or more), the performance cannot be improved unless the Reynolds number is high. This is because the fins are low and hidden in the turbulent boundary layer.

The turbulent boundary layer is composed of a viscous bottom layer (or a laminar bottom layer) that flows in the vicinity of the tube wall in a laminar flow, and a layer (transition zone) intermediate between the laminar and turbulent flows. To compare the thickness of such fin height and viscosity bottom layer, defined as the tube wall of the dimensionless number y ⁺ number 1 below the distance y in the tube center direction.

Here, ρ is the density (kg / m ³ ) of the fluid flowing in the pipe, μ is the viscosity (Pas), u ^* is the friction speed, and is determined by the following equation (2).

In general, the viscous bottom layer is in the range of 0 ≦ y ⁺ ≦ 5.

Table 3 shows the result of calculating y ⁺ corresponding to the fin height in Reference Examples 1 to 3 at the Reynolds number of 4000, which is the boundary between the turbulent region and the transition region. In Table 3, Hf / ID shown in Table 2 is also shown.

From FIG. 11, it can be seen from Table 3 that the performance of Reference Example 2 is improved at a Reynolds number of 4000 or more, and at this time y ⁺ is twice or more that of the viscous bottom layer.

FIG. 15 shows the relationship between the Reynolds number in the turbulent flow area and the value δ ^{* obtained} by dividing the thickness of the viscous bottom layer in the smooth tube (y when y ⁺ = 5) by the maximum inner diameter ID.

  From FIG. 11, the performance of Reference Example 1 starts to improve at Reynolds number of 6000 to 7000. From FIG. 15, the thickness of the viscous bottom layer at Reynolds number of 6000 to 7000 is 0.012.

  From Table 2, the Hf / ID of Reference Example 1 is 0.027, and the performance is improved when the thickness is 0.01 times or more of the viscous bottom layer as in Reference Example 2.

From equation (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 friction speed u ^* is proportional to the friction stress τw in the tube wall. Further, the relationship between the friction stress τw and the pressure loss ΔP (= P1−P2) in the section L is expressed by the following equation (3).

  Darcy-Weisbach equation (Equation 4)

Is substituted into equation (3),

If the fluid temperature and flow velocity are the same, the frictional stress τw is proportional to the pipe friction coefficient f.

FIG. 16 shows a relationship between a value δ ^{* obtained} by dividing the multiple k of the smooth pipe by the maximum inner diameter ID and the multiple k of this and the thickness of the viscous bottom layer (y when y ⁺ = 5).

  The range of the Reynolds number of the heat pump water heater is about 1000 to 7000, and is 4000 to 7000 in the turbulent flow region, so the case of Reynolds numbers 4000 and 7000 is shown.

When the Reynolds number is 4000 and 7000, it is formulated (in the case of 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 of the corrugated tube and the smooth tube ratio (FIG. 10) increase linearly when 0.04 <Hc / OD, so the relational expression between the pressure loss ratio k and Hc / OD is It becomes like this.

Since the performance is improved when Hf / ID is twice or more of the viscous bottom layer δ ^* , the lower limit value of Hf / ID that can improve the performance in the corrugated pipe with an inner groove is expressed by Equation 7.

  Desirably, Equation 8 is obtained.

  In summary, the number 9

  In this case, even if the performance cannot be improved in the Reynolds number range of the heat pump type hot water heater as in the case of the inner surface grooved tube of Reference Example 1, by making this an inner surface grooved corrugated tube, the viscous bottom layer is obtained by the corrugating stirring effect. And the performance can be improved in the Reynolds number range of the heat pump type water heater.

  Table 4 shows ratios when the specifications of Examples 2 to 5 and the heat transfer performance of the inner smooth corrugated pipe (Comparative Example 2) are set to 1.

  In Examples 4 and 5, the upper limit value of Hc / OD (Hc / OD = 0.1) is set so that Examples 2 and 3 have the lower limit value of Hc / OD (Hc / OD = 0.04). Thus, in Examples 2 and 4, Examples 3 and 5 have Hf / ID upper limit values (0 and 0, 0.014, 0.011). 0.037), an internally grooved corrugated tube was manufactured.

  The evaluation of the heat transfer performance in Examples 2 to 5 is expressed as a ratio when the heat transfer performance of Comparative Example 2 is set to 1. When the Reynolds number is 1000, the heat transfer performance equivalent to that of Comparative Example 2 is obtained. It can be seen that, as the Reynolds number increases, the heat transfer performance is improved by several percent compared to Comparative Example 2.

It is explanatory drawing which shows the structure of the heat exchanger tube which concerns on the 1st Embodiment of this invention, (a) shows a general view, (b) is the elements on larger scale of (a), (c) is (a ) Is an enlarged sectional view of an ellipse A. It is explanatory drawing which shows the structure of the heat exchanger tube which concerns on the 2nd Embodiment of this invention. It is explanatory drawing which shows the structure of the heat exchanger which concerns on the 3rd Embodiment of this invention. It is explanatory drawing which shows the structure of the heat exchanger which concerns on the 4th Embodiment of this invention. It is explanatory drawing which shows the structure of the heat exchanger which concerns on the 5th Embodiment of this invention. It is explanatory drawing which shows the structure of the heat exchanger which concerns on the 6th Embodiment of this invention. It is explanatory drawing which shows the structure of the heat exchanger which concerns on the 7th Embodiment of this invention. The relationship between the Hc / OD of the corrugated tube and the heat transfer performance (heat transfer performance ratio with respect to the smooth tube) when the Reynolds number Re is 1000 is shown. The relationship between the twist angle of the corrugated tube and the heat transfer performance when the Reynolds number Re is 1000 (heat transfer performance ratio with respect to the smooth tube) is shown. The relationship between the Hc / OD of the corrugated pipe and the pipe friction coefficient when the Reynolds number Re is 1000 (the pipe friction coefficient ratio with respect to the smooth pipe) is shown. It is a heat-transfer performance measurement result of the spiral inner surface grooved tube of Reference Examples 1-3 and the smooth tube of Comparative Example 1. It is a pressure loss measurement result of the reference examples 1-3 and the comparative example 3. FIG. It is a heat-transfer performance measurement result of a smooth pipe (comparative example 1), an internal smooth corrugated pipe (comparative example 2), and an internal grooved corrugated pipe (comparative example 3, Example 1). It is a pressure-loss measurement result of a smooth pipe (comparative example 1), an internal smooth corrugated pipe (comparative example 2), and an internal grooved corrugated pipe (comparative example 3, Example 1). The relationship between the Reynolds number in the turbulent flow region and the value δ ^{* obtained} by dividing the thickness of the viscous bottom layer in the smooth tube (y when y ⁺ = 5) by the maximum inner diameter ID is shown. This shows the relationship between a multiple of the tube friction coefficient of the smooth tube and the value δ ^{* obtained} by dividing the thickness of the viscous bottom layer (y when y ⁺ = 5) by the maximum inner diameter ID.

Explanation of symbols

1 Corrugated groove 2 Internal groove 10 Corrugated pipe with internal groove

Claims (3)

A corrugated heat transfer tube used as a water tube constituting a heat exchanger,
A spiral inner groove is formed on the inner surface of the corrugated tube,
When the corrugated groove depth of the corrugated shape is Hc and the corrugated outer diameter is OD,
0.04 ≦ Hc / OD ≦ 0.1 is satisfied, and
When the fin height of the inner surface groove is Hf and the maximum inner diameter of the tube is ID,
0.022 {30.7 × (Hc / OD) +1.13} ^(−0.5) ≦ Hf / ID ≦ 0.037
A heat exchanger tube for a heat exchanger characterized by satisfying A corrugated heat transfer tube used as a water tube constituting a heat exchanger,
A spiral inner groove is formed on the inner surface of the corrugated tube,
When the corrugated groove depth of the corrugated shape is Hc and the corrugated outer diameter is OD,
0.04 ≦ Hc / OD ≦ 0.1 is satisfied, and
When the fin height of the inner surface groove is Hf and the maximum inner diameter of the tube is ID,
0.036 {30.7 × (Hc / OD) +1.13} ^(-0.5) ≦ Hf / ID ≦ 0.037
A heat exchanger tube for a heat exchanger characterized by satisfying   A heat exchanger comprising the heat exchanger tube for a heat exchanger according to claim 1.

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