JPS63172894A - Heat exchanger - Google Patents
Heat exchangerInfo
- Publication number
- JPS63172894A JPS63172894A JP450287A JP450287A JPS63172894A JP S63172894 A JPS63172894 A JP S63172894A JP 450287 A JP450287 A JP 450287A JP 450287 A JP450287 A JP 450287A JP S63172894 A JPS63172894 A JP S63172894A
- Authority
- JP
- Japan
- Prior art keywords
- groove
- pipe
- heat transfer
- spiral
- heat exchanger
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Granted
Links
- 239000012530 fluid Substances 0.000 claims abstract description 9
- 239000007788 liquid Substances 0.000 abstract description 26
- 239000002826 coolant Substances 0.000 abstract 5
- 239000003507 refrigerant Substances 0.000 description 37
- 230000000694 effects Effects 0.000 description 10
- 238000001704 evaporation Methods 0.000 description 5
- 230000008020 evaporation Effects 0.000 description 5
- 238000009833 condensation Methods 0.000 description 4
- 230000005494 condensation Effects 0.000 description 4
- 238000000034 method Methods 0.000 description 2
- 230000008569 process Effects 0.000 description 2
- 230000007704 transition Effects 0.000 description 2
- VOPWNXZWBYDODV-UHFFFAOYSA-N Chlorodifluoromethane Chemical compound FC(F)Cl VOPWNXZWBYDODV-UHFFFAOYSA-N 0.000 description 1
- 241000700124 Octodon degus Species 0.000 description 1
- 230000009471 action Effects 0.000 description 1
- 230000008859 change Effects 0.000 description 1
- 239000003795 chemical substances by application Substances 0.000 description 1
- 230000007423 decrease Effects 0.000 description 1
- 238000001035 drying Methods 0.000 description 1
- 238000000227 grinding Methods 0.000 description 1
- 238000004519 manufacturing process Methods 0.000 description 1
- 230000007246 mechanism Effects 0.000 description 1
- 230000001737 promoting effect Effects 0.000 description 1
- 238000010992 reflux Methods 0.000 description 1
- 238000007665 sagging Methods 0.000 description 1
- 239000007921 spray Substances 0.000 description 1
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F28—HEAT EXCHANGE IN GENERAL
- F28F—DETAILS OF HEAT-EXCHANGE AND HEAT-TRANSFER APPARATUS, OF GENERAL APPLICATION
- F28F1/00—Tubular elements; Assemblies of tubular elements
- F28F1/10—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses
- F28F1/40—Tubular elements and assemblies thereof with means for increasing heat-transfer area, e.g. with fins, with projections, with recesses the means being only inside the tubular element
Landscapes
- Physics & Mathematics (AREA)
- Engineering & Computer Science (AREA)
- Geometry (AREA)
- Thermal Sciences (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Heat-Exchange Devices With Radiators And Conduit Assemblies (AREA)
Abstract
Description
【発明の詳細な説明】
産業上の利用分野
本発明は空気調和機や冷凍機等に使用され、冷媒と空気
等の流体間で熱の授受を行う熱交換器に関するものであ
る。DETAILED DESCRIPTION OF THE INVENTION Field of the Invention The present invention relates to a heat exchanger used in air conditioners, refrigerators, etc., which transfers heat between a refrigerant and a fluid such as air.
従来の技術
冷媒等の作動流体が伝熱管内を相変化しながら流動する
熱交換器としては、従来、第８図に示すようなフィン付
熱交換器が用いられている。これは、一定間隔で平行に
並べられたフィン１とフィン１に直角に挿通された伝熱
管２とからなり、伝熱管２内を流れる冷媒とフィン１間
を流れる空気との間で熱交換が行われる。図中矢印３は
冷媒の流動方向を示す。伝熱管２としては、第９図に示
すように、管内壁にらせん溝４が設けられており、らせ
ん溝４の管軸に対する溝ねじれ角β及びらせん溝５の溝
深さｈは一定に保たれた管内らせん溝付管が用いられて
いる。BACKGROUND OF THE INVENTION Conventionally, a finned heat exchanger as shown in FIG. 8 has been used as a heat exchanger in which a working fluid such as a refrigerant flows through heat transfer tubes while undergoing a phase change. It consists of fins 1 arranged in parallel at regular intervals and heat exchanger tubes 2 inserted at right angles to the fins 1. Heat exchange occurs between the refrigerant flowing inside the heat exchanger tubes 2 and the air flowing between the fins 1. It will be done. Arrow 3 in the figure indicates the flow direction of the refrigerant. As shown in FIG. 9, the heat transfer tube 2 is provided with a helical groove 4 on the inner wall of the tube, and the groove torsion angle β of the helical groove 4 with respect to the tube axis and the groove depth h of the helical groove 5 are kept constant. A sagging internal spiral grooved tube is used.
この管内らせん溝付管はらせん溝５を設けることによっ
て、蒸発熱伝達の場合は水平ならせん溝付管内の管底部
を流れている液冷媒が毛細現象によって溝内を引き上げ
られ、管内伝熱面に形成される冷媒液膜の平均厚さが薄
くなり、凝縮熱伝達の場合は凝縮液が表面張力の作用で
らせん溝４底部に集まり、管内伝熱面に形成される凝縮
液膜の平均厚さが薄くなり、蒸発時、凝縮時共、管内伝
熱性能が向上すると言われていた。By providing the helical groove 5 in this internal spiral grooved tube, in the case of evaporative heat transfer, the liquid refrigerant flowing at the bottom of the horizontal spiral grooved tube is pulled up in the groove by capillary phenomenon, and the internal heat transfer surface In the case of condensation heat transfer, the condensate gathers at the bottom of the spiral groove 4 due to surface tension, and the average thickness of the condensate film formed on the heat transfer surface inside the tube decreases. It was said that this would improve the heat transfer performance within the tube during both evaporation and condensation.
発明が解決しようとする問題点
しかし、例えば、蒸発熱伝達の場合に管内伝熱面の全て
に薄い厚さの冷媒液膜が形成され、著しい伝熱促進効果
が得られるのは、冷媒の乾き度が大きい領域、すなわち
蒸発過程の後期だけである。Problems to be Solved by the Invention However, for example, in the case of evaporative heat transfer, a thin refrigerant liquid film is formed on all the heat transfer surfaces in the tube, and the reason why a remarkable heat transfer promotion effect is obtained is due to the drying of the refrigerant. It is only in the region where the temperature is large, that is, in the later stages of the evaporation process.
一方、蒸発過程の初期においては、冷媒の乾き度は小さ
く、管内の液冷媒の流量が多いために、液冷媒によって
らせん溝は埋もれてしまい、液冷媒はらせん溝の上を通
過してしまう。従って、上述した様な蒸発メカニズムに
よる著しい伝熱促進効果は望めない。このことは、凝縮
熱伝達の場合についても同様に考えられる。従って、管
内冷媒が低乾き度の場合に、管内熱伝達率が低いために
、熱交換器としての性能（熱通過率）が低いという問題
を有していた。On the other hand, at the beginning of the evaporation process, the degree of dryness of the refrigerant is low and the flow rate of the liquid refrigerant in the tube is large, so the helical groove is buried by the liquid refrigerant and the liquid refrigerant passes over the helical groove. Therefore, a significant heat transfer promoting effect due to the evaporation mechanism as described above cannot be expected. This can be similarly considered in the case of condensation heat transfer. Therefore, when the refrigerant in the tubes has a low degree of dryness, the heat transfer coefficient in the tubes is low, resulting in a problem that the performance as a heat exchanger (heat transfer rate) is low.
そこで、本発明は管内らせん溝付管内のらせん溝を工夫
することによって、冷媒の低乾き度域から中乾き度域に
おける管内伝達率を向上させ、かつ、高乾き度域におけ
る圧力損失を低減することによって、高性能な熱交換器
を得ようとするものである。Therefore, the present invention improves the intra-pipe transfer coefficient in the low to medium dryness range of the refrigerant and reduces the pressure loss in the high dryness range by devising the spiral groove in the pipe. By doing so, the aim is to obtain a high-performance heat exchanger.
問題点を解決するための手段
上記問題点を解決する本発明の技術的手段は、熱交換器
を構成する伝熱管内に、らせん溝の管軸に対する溝ねじ
れ角及びらせん溝の溝深さがそれぞれ異なる少なくとも
２種類のらせん溝を設け、かつ、管内流体の高乾き度域
に比して、低乾き度域において溝ねじれ角及び溝深さを
それぞれより大きくするものである。Means for Solving the Problems The technical means of the present invention for solving the above-mentioned problems is that the helical grooves have a groove helix angle with respect to the tube axis and a groove depth of the helical grooves in the heat exchanger tubes constituting the heat exchanger. At least two different types of helical grooves are provided, and the groove torsion angle and groove depth are made larger in the low dryness region than in the high dryness region of the fluid in the pipe.
作 用 この技術的手段による作用は次のようになる。For production The effect of this technical means is as follows.
すなわち、上述したように、熱交換器を構成する、水平
に設置された伝熱管内に、らせん溝の管軸に対する溝ね
じれ角β及びらせん溝の溝深さｈが異なる少なくとも２
種類のらせん溝を設け、かつ、管内流体の低乾き度域に
おける溝ねじれ角をβ 、溝深さをｈｌ とし、高乾
き度域における溝ねじれ角をβ２．溝深さをｈ２ とす
ると、β１〉β２及びｈｌ〉ｈ２なる関係があるため、
本発明による熱交換器を蒸発器として使用する場合には
、低乾き度域から中乾き度域（例えば、冷媒Ｒ−２２．
管径４鵡の場合、乾き度Ｉ≦０．７）、すなわち、管内
の液冷媒の流量が多い領域で、溝深さｈｌ が大きいた
めに液冷媒によって埋もれることが少なく、かつ、溝ね
じれ角β１が大きいために比較的流速の小さい液冷媒が
らせん溝に沿って流れやすくなる。例えば、流動状態は
波状流から環状流へ遷移しやすくなり、管内の全らせん
溝の山頂部分に均一な厚さの冷媒液膜が形成され伝熱性
能を大幅に向上させることができる。また、高乾き度域
（ｘ２０．７）、すなわち、液冷媒の流量が少ない領域
では、ある程度の溝深さく例えば管径４鰭では０．１〜
○＊２ａ）’ｒあればらせん溝内を液冷媒が毛細管現象
によって引き上げられるために低乾き度域はど溝深さｈ
２を大きくしなくても、流動状態は還伏流あるいは環状
噴霧流に遷移しやすくなり、管内の全溝面上に薄い厚さ
の冷媒液膜が形成され、高い熱伝達率が得られる。更に
、溝ねじれ角β２カ砥乾き度より小さいため圧力損失の
低減が図れる。これは、溝ねじれ角βが管内熱伝達率に
与える影響は小さいにも拘らず、圧力損失に与える影響
が大きいからである。すなわち、溝ねじれ角βが大きく
なるほど圧力損失は増加する。That is, as described above, in a horizontally installed heat transfer tube constituting a heat exchanger, there are at least two helical grooves in which the groove torsion angle β with respect to the tube axis of the helical groove and the groove depth h of the helical groove are different.
Types of spiral grooves are provided, and the groove torsion angle in the low dryness region of the fluid in the pipe is β, the groove depth is hl, and the groove torsion angle in the high dryness region is β2. If the groove depth is h2, there are relationships β1>β2 and hl>h2, so
When the heat exchanger according to the present invention is used as an evaporator, it may be used in a low to medium dryness range (for example, refrigerant R-22.
In the case of a pipe diameter of 4 mm, in the dryness I≦0.7), that is, in the region where the flow rate of liquid refrigerant in the pipe is large, the groove depth hl is large, so it is less likely to be buried by liquid refrigerant, and the groove torsion angle is Since β1 is large, liquid refrigerant having a relatively low flow rate easily flows along the spiral groove. For example, the flow state can easily transition from a wavy flow to an annular flow, and a refrigerant liquid film with a uniform thickness is formed on the crests of all the spiral grooves in the pipe, making it possible to significantly improve heat transfer performance. In addition, in the high dryness region (x20.7), that is, in the region where the flow rate of liquid refrigerant is small, the groove depth is to a certain extent, for example, 0.1~
○*2a) If 'r, the liquid refrigerant is pulled up in the spiral groove by capillary phenomenon, so the groove depth h in the low dryness area
Even if 2 is not increased, the flow state easily transitions to a reflux flow or an annular spray flow, a thin refrigerant liquid film is formed on all groove surfaces in the tube, and a high heat transfer coefficient can be obtained. Furthermore, since the groove helix angle β2 is smaller than the grinding dryness, pressure loss can be reduced. This is because although the groove torsion angle β has a small effect on the in-pipe heat transfer coefficient, it has a large effect on the pressure loss. That is, the pressure loss increases as the groove torsion angle β increases.
また、本発明は凝縮熱伝達においても、蒸発熱伝達の場
合と同様の効果を発揮する。Moreover, the present invention exhibits the same effect in condensation heat transfer as in the case of evaporative heat transfer.
以上より、冷媒低乾き度域から中乾き度域における管内
熱伝達率の向上、及び高乾き度域における管内圧力損失
の低減が図れ、すなわち、管内伝熱性能の向上が図れ、
従って、熱交換器の高性能化が可能になる。From the above, it is possible to improve the heat transfer coefficient in the pipe in the low to medium dryness range of the refrigerant, and to reduce the pressure loss in the pipe in the high dryness range, that is, to improve the heat transfer performance in the pipe.
Therefore, it becomes possible to improve the performance of the heat exchanger.
実施例
以下、本発明の一実施例を第１図〜第７図に基づいて説
明する。EXAMPLE Hereinafter, an example of the present invention will be described based on FIGS. 1 to 7.
第１図は本発明の一実施例の熱交換器の一部欠截正面図
である。１０は一定間隔に並設したフィンであり、その
フィン１０を貫通して配列した伝熱管１１、各伝熱管１
１を互いに結合するＵ字形ベンド１２、及びエンドプレ
ート１３よす熱交換器が構成されている。伝熱管１１内
を矢印１４方向に冷媒が流動し、その冷媒と、フィン１
ｏ間を流れる空気との間で熱交換が行われる。そして、
各伝熱管１１ａ、１１ｂ、１１ｃ内壁には、それぞれ第
２図〜第７図に示すような管軸に対する垂直断面形状が
三角状のらせん溝１５ａ、１５ｂ、１５ｃが刻設されて
いる。すなわち、冷媒乾き度を低。FIG. 1 is a partially cutaway front view of a heat exchanger according to an embodiment of the present invention. Reference numeral 10 denotes fins arranged in parallel at regular intervals, heat exchanger tubes 11 arranged through the fins 10, and each heat exchanger tube 1
A U-shaped bend 12 connecting the parts 1 to each other and an end plate 13 constitute a heat exchanger. A refrigerant flows in the direction of the arrow 14 inside the heat transfer tube 11, and the refrigerant and the fins 1
Heat exchange takes place with the air flowing between the two. and,
Spiral grooves 15a, 15b, and 15c each having a triangular cross-sectional shape perpendicular to the tube axis as shown in FIGS. 2 to 7 are cut into the inner walls of each of the heat exchanger tubes 11a, 11b, and 11c. In other words, the dryness of the refrigerant is low.
中、高と三つに分け、低乾き度域の伝熱管１１ａのらせ
ん溝１５ａの管軸に対する溝ねじれ角をβ１゜溝深さを
ｈｌ、同様に、中乾き度域の伝熱管１１ｂのらせん溝１
６ｂの溝ねじれ角をβ２．溝深さをｈ２゜高乾き度域の
伝熱管１１ｃのらせん溝１６Ｃの溝ねじれ角をβ 、溝
深さをｈ３とするとβ１〉β２〉β３．ｈｌ〉ｈ２＞ｈ
３なる関係がある。Divided into three parts, medium and high, the helical groove 15a of the heat exchanger tube 11a in the low dryness area has a groove twist angle of β1° with respect to the tube axis, and the groove depth is hl, and similarly, the helical groove of the heat exchanger tube 11b in the medium dryness area Groove 1
The groove twist angle of 6b is β2. Let the groove depth be h2°, the groove twist angle of the spiral groove 16C of the heat exchanger tube 11c in the high dryness region be β, and the groove depth be h3, then β1>β2>β3. hl>h2>h
There are three relationships.
次に、この一実施例の構成における作用を説明する。Next, the operation of the configuration of this embodiment will be explained.
本発明による一実施例の熱交換器を蒸発器として使用す
る場合、低乾き度域、すなわち、管内の液冷媒流量が多
い領域では、流動状態は層状流あるいはスラグ流である
ことが多いが溝ねじれ角β１及び溝深さｈｌがβ２．β
３及びｈ２．ｈ３よシ大きいために第３図に示すように
、水平管の下部には液冷媒１６ａがたまっているが、上
部へ液冷媒１８ａが表面張力によりらせん溝１５ａに沿
って流れやすくなって、上部では比較的薄い厚みの冷媒
液膜１７ａが形成される。従って、伝熱管１１ａの管内
上部における蒸発熱伝達率が向上する。次に、中乾き度
域、すなわち、流動状態が通常波状流から半環状流のよ
うな状態では、比較的液冷媒流量が少なく、らせん溝１
５ｂ内を液冷媒１６ｂが毛細管現象によって引き上げら
れるが、溝深さｈ２は伝熱管１１ｂ内下部において液冷
媒１６ｂがたまっても三角状山頂部１８が完全に埋まら
ない高さであるので（ｈ２〈ｈｌ）、管内上部はもちろ
ん、下部においても比較的均一な厚さの冷媒液膜１７ｂ
が形成され、伝熱管１１ｂ内において流動状態が環状流
に遷移しやすくなる。従って、伝熱管１１ｂ内における
蒸発熱伝達率が大幅に向上すると共に、溝ねじれ角β２
はβ２〈β、であるので、圧力損失の増大は抑えられる
。更に、高乾き度域においては、液冷媒流量もかなり少
なくなシ、気相によって液冷媒が加速されるため、伝熱
管１１Ｃ内のらせん溝１５ｃ上に形成される冷媒液膜１
７ｃはかなり薄くなり、従って、蒸発熱伝達率は最大と
なる。この場合、溝深さｈ３は冷媒液膜１７ｃの厚みが
薄くなることより、ｈ３〈ｈ２としており、溝ねじれ角
β３もβ３〈β２としている。これは、高乾き度域にお
いては、溝深さｈ３及び溝ねじれ角β３が蒸発熱伝達率
に与える影響は小さく、逆に、圧力損失への影響が大き
いため、すなわち、ｈ３及びβ３が大きい程、圧力損失
も大きくなることを考慮しているからである。When the heat exchanger according to one embodiment of the present invention is used as an evaporator, in a low dryness region, that is, in a region where the flow rate of liquid refrigerant in the pipes is large, the flow state is often a laminar flow or a slug flow. The helix angle β1 and the groove depth hl are β2. β
3 and h2. As shown in Fig. 3, the liquid refrigerant 16a accumulates at the bottom of the horizontal pipe because it is larger than h3. In this case, a relatively thin refrigerant liquid film 17a is formed. Therefore, the evaporative heat transfer coefficient in the upper part of the heat exchanger tube 11a is improved. Next, in the medium dryness region, that is, when the flow state is from a normal wavy flow to a semi-annular flow, the liquid refrigerant flow rate is relatively small, and the spiral groove 1
Although the liquid refrigerant 16b is pulled up inside the heat transfer tube 11b by capillary action, the groove depth h2 is such that even if the liquid refrigerant 16b accumulates in the lower part of the heat transfer tube 11b, the triangular peak 18 is not completely buried (h2< hl), the refrigerant liquid film 17b has a relatively uniform thickness not only in the upper part of the pipe but also in the lower part.
is formed, and the flow state easily changes to an annular flow within the heat exchanger tube 11b. Therefore, the evaporative heat transfer coefficient within the heat exchanger tube 11b is significantly improved, and the groove torsion angle β2
Since β2<β, the increase in pressure loss can be suppressed. Furthermore, in a high dryness region, the flow rate of liquid refrigerant is quite small, and the liquid refrigerant is accelerated by the gas phase, so that the refrigerant liquid film 1 formed on the spiral groove 15c in the heat transfer tube 11C
7c becomes much thinner and therefore the evaporative heat transfer coefficient is maximum. In this case, since the thickness of the refrigerant liquid film 17c becomes thinner, the groove depth h3 is set as h3<h2, and the groove torsion angle β3 is also set as β3<β2. This is because, in a high dryness region, the groove depth h3 and the groove torsion angle β3 have a small effect on the evaporative heat transfer coefficient, but conversely, they have a large effect on the pressure loss, that is, the larger h3 and β3 are, the more This is because it takes into consideration that the pressure loss will also increase.
以上より、冷媒低乾き度域から中乾き度域における蒸発
熱伝達率の向上及び、高乾き度域における管内圧力損失
の低減が図れ、蒸発管としての伝熱性能が向上する。As described above, it is possible to improve the evaporative heat transfer coefficient in the low to medium dryness region of the refrigerant and reduce the pressure loss inside the pipe in the high dryness region, and improve the heat transfer performance as an evaporation tube.
また、本発明は、凝縮器として用いた場合においても蒸
発器の場合と同様の効果を発揮する。更に、上記実施例
では、管内らせん溝１ｓａ、１ｓｂ。Furthermore, the present invention exhibits the same effects as when used as an evaporator even when used as a condenser. Furthermore, in the above embodiment, the internal spiral grooves 1sa and 1sb.
１５ｃは断面が三角状溝であるが、断面形状が三角形以
外の多′角形であっても同様の効果が得られる。Although the groove 15c has a triangular cross section, the same effect can be obtained even if the cross section is a polygonal shape other than a triangle.
発明の効果
以上のように本発明は、内部を相変化する流体が流動し
、管内壁にらせん溝の刻設されている伝熱管とから熱交
換器を構成し、伝熱管内に、らせん溝の管軸に対する溝
ねじれ角及び溝深さがそれぞれ異なる少なくとも２種類
のらせん溝を設け、かつ、管内流体の高乾き度域に比し
て、低乾き度域において、溝ねじれ角及び溝深さをそれ
ぞれより大きくしているために、冷媒代乾き度域から中
乾き度域における管内熱伝達率の大幅な向上、及び、高
乾き度域における管内圧力損失の低減が図れ、熱交換器
としての伝熱性能が向上する。Effects of the Invention As described above, the present invention comprises a heat exchanger including a heat exchanger tube in which a phase-changing fluid flows, and a spiral groove is carved in the inner wall of the tube. At least two types of helical grooves having different groove torsion angles and groove depths with respect to the pipe axis are provided, and the groove torsion angle and groove depth are different in a low dryness region than in a high dryness region of the fluid in the pipe. Since these are made larger, it is possible to significantly improve the heat transfer coefficient in the pipe from the refrigerant cost dryness range to the medium dryness range, and to reduce the pressure loss in the pipe in the high dryness range, making it suitable for use as a heat exchanger. Heat transfer performance is improved.
第１図は本発明の一実施例による熱交換器の一部欠截正
面図、第２図は第１図中の伝熱管１１ａの一部欠截正面
図、第３図は第２図のＡ−Ａ線における断面図、第４図
は第１図中の伝熱管１１ｂの一部欠截正面図、第５図は
第４図のＢ−Ｂ線における断面図、第６図は第１図中の
伝熱管１１ｃの一部欠截正面図、第７図は第６図のＣ−
Ｃ線における断面図、第８図は従来の熱交換器の斜視図
、第９図は第８図中の伝熱管の一部欠截正面図である。
１１ 、１１ ａ 、 １ ｌ ｂ 、 １１ ｃ−−
伝熱管、１５ａ、１５ｂ、１５ｃ・・・・−らせん溝、
β１．β２゜β３・・・・・・溝ねじれ角、ｈｌ、ｈ２
．ｈ３・・・・・・溝深さ。
代理人の氏名 弁理士 中 尾 敏 男 ほか１名Ｉｌ
ｌん’１．Ｉ７ｂ、ｌ７Ｄ−’ｉデぐ差へ乍紫１５１１
、Ｉ５ｂ、７５／、−−らセンし１転第１図1 is a partially cutaway front view of a heat exchanger according to an embodiment of the present invention, FIG. 2 is a partially cutaway front view of a heat exchanger tube 11a in FIG. 1, and FIG. 3 is a partially cutaway front view of a heat exchanger 11a in FIG. 4 is a partially cutaway front view of the heat exchanger tube 11b in FIG. 1, FIG. 5 is a sectional view taken along the line BB in FIG. 4, and FIG. A partially cutaway front view of the heat exchanger tube 11c in the figure, FIG. 7 is C- in FIG. 6.
8 is a perspective view of a conventional heat exchanger, and FIG. 9 is a partially cutaway front view of a heat exchanger tube in FIG. 8. 11, 11a, 1lb, 11c--
Heat exchanger tubes, 15a, 15b, 15c...-Spiral grooves,
β1. β2゜β3...Groove torsion angle, hl, h2
．． h3...Groove depth. Name of agent: Patent attorney Toshio Nakao and one other person
ln'1. I7b, l7D-'i Degu difference 1511
, I5b, 75/, -- Rasenshi 1st turn Figure 1
Claims (1)
刻設されている伝熱管から構成され、前記伝熱管内にら
せん溝の管軸に対する溝ねじれ角と、らせん溝の溝深さ
がそれぞれ異なる少なくとも２種類のらせん溝を設け、
かつ、管内流体の高乾き度域に比して低乾き度域にて、
溝ねじれ角と溝深さをそれぞれより大きくした熱交換器
。It consists of a heat transfer tube in which a phase-changing fluid flows, and a spiral groove is carved on the inner wall of the tube, and the groove twist angle with respect to the tube axis of the spiral groove and the groove depth of the spiral groove are determined in the heat transfer tube. At least two different types of helical grooves are provided,
And, in the low dryness area compared to the high dryness area of the fluid in the pipe,
A heat exchanger with larger groove helix angle and groove depth.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP62004502A JPH07109354B2 (en) | 1987-01-12 | 1987-01-12 | Heat exchanger |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP62004502A JPH07109354B2 (en) | 1987-01-12 | 1987-01-12 | Heat exchanger |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPS63172894A true JPS63172894A (en) | 1988-07-16 |
| JPH07109354B2 JPH07109354B2 (en) | 1995-11-22 |
Family
ID=11585832
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP62004502A Expired - Lifetime JPH07109354B2 (en) | 1987-01-12 | 1987-01-12 | Heat exchanger |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPH07109354B2 (en) |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH0425982U (en) * | 1990-06-20 | 1992-03-02 | ||
| WO2001031275A1 (en) * | 1999-10-28 | 2001-05-03 | Mitsubishi Shindoh Co., Ltd. | Heat exchanger and heat exchanging apparatus |
| WO2013094084A1 (en) * | 2011-12-19 | 2013-06-27 | 三菱電機株式会社 | Air conditioner |
Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5938596A (en) * | 1982-08-25 | 1984-03-02 | Matsushita Electric Ind Co Ltd | Heat exchanger |
| JPS61110891A (en) * | 1984-11-01 | 1986-05-29 | Matsushita Electric Ind Co Ltd | Heat exchanger |
| JPS61280393A (en) * | 1985-05-17 | 1986-12-10 | Sanyo Electric Co Ltd | Heat transfer tube |
-
1987
- 1987-01-12 JP JP62004502A patent/JPH07109354B2/en not_active Expired - Lifetime
Patent Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPS5938596A (en) * | 1982-08-25 | 1984-03-02 | Matsushita Electric Ind Co Ltd | Heat exchanger |
| JPS61110891A (en) * | 1984-11-01 | 1986-05-29 | Matsushita Electric Ind Co Ltd | Heat exchanger |
| JPS61280393A (en) * | 1985-05-17 | 1986-12-10 | Sanyo Electric Co Ltd | Heat transfer tube |
Cited By (4)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH0425982U (en) * | 1990-06-20 | 1992-03-02 | ||
| WO2001031275A1 (en) * | 1999-10-28 | 2001-05-03 | Mitsubishi Shindoh Co., Ltd. | Heat exchanger and heat exchanging apparatus |
| WO2013094084A1 (en) * | 2011-12-19 | 2013-06-27 | 三菱電機株式会社 | Air conditioner |
| US9506700B2 (en) | 2011-12-19 | 2016-11-29 | Mitsubishi Electric Corporation | Air-conditioning apparatus |
Also Published As
| Publication number | Publication date |
|---|---|
| JPH07109354B2 (en) | 1995-11-22 |
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