JP2017129360A - Heat exchanger - Google Patents
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本発明は、フィンを備えず冷媒通路を有する伝熱プレートで熱交換用コア部を形成した熱交換器に関する。 The present invention relates to a heat exchanger in which a heat exchanging core portion is formed by a heat transfer plate having no refrigerant and having a refrigerant passage.
従来の熱交換器、例えば、車両空調用蒸発器においては、冷媒通路を有する伝熱プレート相互の間に、コルゲートフィン等のフィンを介在したものが一般的であるが、フィンの介在により、圧損が増大し、コスト増加、大型化する等の難点がある。 In conventional heat exchangers, for example, vehicle air conditioner evaporators, fins such as corrugated fins are generally interposed between heat transfer plates having refrigerant passages. However, there are problems such as an increase in cost and an increase in size.
このため、特許文献１および特許文献２では、フィンを有さず冷媒通路を有した伝熱プレート（フィンレス伝熱プレート）のみで熱交換用コア部を形成した熱交換器を提案している。 For this reason, Patent Document 1 and Patent Document 2 propose a heat exchanger in which a heat exchanging core portion is formed only by a heat transfer plate (finless heat transfer plate) having no refrigerant and having a refrigerant passage.
フィンレス伝熱プレートでは、熱交換面積を大きく得ることが難しいため、熱交換性能を確保しにくい。
フィンレス伝熱プレートを用いた熱交換器における熱交換性能を高めるため、特許文献１では、伝熱プレートに形成される凸部のピッチ、特許文献２では、凸部と対向する凹部との隙間をそれぞれ数値限定している。
In the finless heat transfer plate, it is difficult to obtain a large heat exchange area, so it is difficult to ensure heat exchange performance.
In order to enhance the heat exchange performance in a heat exchanger using a finless heat transfer plate, in Patent Document 1, the pitch of the convex portions formed on the heat transfer plate is defined, and in Patent Document 2, the gap between the convex portions and the concave portions facing each other is defined. Each number is limited.
しかしながら、これら伝熱プレートの凸部のピッチ及び、凸部と対向する凹部との隙間では、熱交換器の熱交換性能を高精度に推定できるパラメータとは言い難く、熱交換性能の向上には限界があった。 However, the pitch between the convex portions of the heat transfer plates and the gap between the convex portions and the concave portions facing the convex portions are hardly parameters that can accurately estimate the heat exchange performance of the heat exchanger. There was a limit.
本発明は、このような従来の課題に着目してなされたもので、フィンレス伝熱プレートを用いた熱交換器の熱交換性能を十分に高めることを目的とする。 The present invention has been made paying attention to such a conventional problem, and an object thereof is to sufficiently enhance the heat exchange performance of a heat exchanger using a finless heat transfer plate.
かかる課題を解決するため第１の発明にかかる熱交換器は、
複数枚の伝熱プレートを間隔を開けて平行に配設し、各伝熱プレートには表裏両面に、複数の台形状の凸部を一方向に並べて配設し、各凸部内に全長にわたって冷媒通路を形成し、
前記各伝熱プレートの凸部が、対向する伝熱プレートの凸部間の凹部に対向するように配設し、
前記各伝熱プレート相互間に、前記一方向と平行な方向に流体を流通させるように積層して構成した熱交換用コア部を備えた熱交換器において、
前記伝熱プレートの凸部の流体流通方向上流側の角部と、該伝熱プレートに対向する伝熱プレートの前記凸部の角部に近接する凸部の流体流通方向下流側の角部との間を流通する流体流の主軸が前記対向する伝熱プレート相互を結ぶ垂直方向に対する鋭角側のなす角θを、０．２５〜０．６５の範囲内の値に設定したことを特徴とする。
In order to solve this problem, the heat exchanger according to the first invention is:
A plurality of heat transfer plates are arranged in parallel at intervals, and a plurality of trapezoidal convex portions are arranged in one direction on both the front and back surfaces of each heat transfer plate. Form a passage,
The convex portions of the heat transfer plates are disposed so as to face the concave portions between the convex portions of the opposing heat transfer plates,
Between the heat transfer plates, a heat exchanger including a heat exchanging core portion configured to be laminated so that a fluid flows in a direction parallel to the one direction,
A corner on the upstream side in the fluid flow direction of the convex portion of the heat transfer plate, and a corner on the downstream side in the fluid flow direction of the convex portion close to the corner of the convex portion of the heat transfer plate facing the heat transfer plate; An angle θ formed by an acute angle side with respect to a vertical direction in which a main axis of a fluid flow flowing between the two is opposite to each other is set to a value within a range of 0.25 to 0.65. .
また、第２の発明に係る熱交換器は、
上記同様の熱交換器において、
前記伝熱プレートの凸部の流体流通方向上流側の角部と、該伝熱プレートに対向する伝熱プレートの前記凸部の角部に近接する凸部の角部との距離ｄｃと、前記対向する各伝熱プレートの前記各角部の流体流通方向上流側に隣接して対向する面相互の距離ｄ１とで定義される縮流比β（＝ｄｃ／ｄ１）が、０．１４〜０．４２の範囲内の値に設定され、
前記縮流比βは、前記伝熱プレートの凸部の基準面からの高さをａｕ、前記基準面と該基準面に対向する伝熱プレートの基準面との距離を２ｈ、前記角部同士の距離ｄｃ方向と前記対向する基準面同士の距離２ｈ方向とのなす角度をα’としたとき、下記の式で算出されることを特徴とする。
In the same heat exchanger as above,
And the heat transfer corner of the fluid flow direction upstream side of the convex portion of the plate, and the distance d c between the convex corner portion adjacent to the corner portion of the convex portion of the heat transfer plate facing the heat transfer plate, The contraction ratio β (= d c / d 1 ) defined by the distance d 1 between the opposing surfaces adjacent to the upstream side in the fluid flow direction of the corners of the opposing heat transfer plates is 0. Set to a value in the range of .14 to 0.42,
The contraction ratio β is the height of the convex portion of the heat transfer plate from the reference surface a u , the distance between the reference surface and the reference surface of the heat transfer plate facing the reference surface is 2 h, and the corner portion when the angle between the distance 2h direction of the reference faces of the distance d c direction between the opposing was alpha ', characterized in that it is calculated by the following equation.
また、第３の発明に係る熱交換器は、
上記同様の熱交換器において、
前記伝熱プレートの凸部の流体流通方向上流側の角部と、該伝熱プレートに対向する伝熱プレートの前記凸部の角部に近接する凸部の角部との距離ｄｃと、前記対向する各伝熱プレートの前記各角部の流体流通方向下流側に隣接して対向する面相互の距離ｄ２とで定義される膨張比ｄ２／ｄｃを、２．０〜４．２の範囲内の値に設定したことを特徴とする。
The heat exchanger according to the third invention is
In the same heat exchanger as above,
And the heat transfer corner of the fluid flow direction upstream side of the convex portion of the plate, and the distance d c between the convex corner portion adjacent to the corner portion of the convex portion of the heat transfer plate facing the heat transfer plate, the expansion ratio d 2 / d c which is adjacent to the fluid flow direction downstream side is defined by the distance d 2 opposing faces each other each corner of the heat transfer plate to the opposed, from 2.0 to 4. It is characterized by being set to a value within a range of 2.
第１〜第３の発明にかかる熱交換器及びこれら熱交換器に用いられる伝熱プレートによれば、フィンレス伝熱プレートにおいて、なす角θ、縮流比ｄｃ／ｄ１、膨張比ｄ２／ｄｃを各数値範囲内に設定することにより、高い熱交換性能を得られる熱交換器を容易に形成することができる。 According to the heat exchangers according to the first to third inventions and the heat transfer plates used in these heat exchangers, in the finless heat transfer plate, the angle θ, the contraction ratio d c / d 1 , and the expansion ratio d 2. / a d c by setting in each numerical range, it is possible to easily form the heat exchanger obtain high heat exchange performance.
以下に本発明の実施の形態を説明する。
図１及び図２は、本発明の実施形態に係る熱交換器１で、例えば、ヒートポンプ式の車両用空調装置の蒸発器に適用されるが、これに限らず凝縮器を含めた任意の熱交換器に適用できる。
Embodiments of the present invention will be described below.
1 and 2 show a heat exchanger 1 according to an embodiment of the present invention. For example, the heat exchanger 1 is applied to an evaporator of a heat pump type vehicle air conditioner. Applicable to exchangers.
熱交換器１は、複数枚の伝熱プレート２を間隔ｃを開けて平行に配設して熱交換用コア部３が形成され、該熱交換用コア部３の上下端部にヘッダ４，５を連結した概略構成を有している。 The heat exchanger 1 includes a plurality of heat transfer plates 2 arranged in parallel with a gap c therebetween to form a heat exchanging core portion 3, and headers 4 on the upper and lower ends of the heat exchanging core portion 3. 5 is connected.
図３（図２のＡ−Ａ断面）に示すように、各伝熱プレート２には表裏両面にそれぞれ上下方向に延びる複数の台形状の凸部２１が、水平方向に列設され、各凸部２１内に全長にわたって冷媒通路２３が形成されている。 As shown in FIG. 3 (cross section AA in FIG. 2), each heat transfer plate 2 is provided with a plurality of trapezoidal convex portions 21 extending in the vertical direction on both the front and back surfaces in a horizontal direction. A refrigerant passage 23 is formed in the portion 21 over the entire length.
各伝熱プレート２の凸部２１は、対向する伝熱プレートの凸部２１間の凹部２２に対向するように配設されている。
ヘッダ４，５には、各伝熱プレート２の上端部及び下端部が挿入されて、各冷媒通路２３は、ヘッダ４，５内部空間（冷媒通路）と連通している。
The convex portion 21 of each heat transfer plate 2 is disposed so as to face the concave portion 22 between the convex portions 21 of the opposing heat transfer plates.
The headers 4 and 5 are inserted with the upper end portion and the lower end portion of each heat transfer plate 2, and each refrigerant passage 23 communicates with the internal space (refrigerant passage) of the headers 4 and 5.
ヘッダ４には、中央部とその両側の３つの空間を仕切る２個の隔壁４１が形成され、ヘッダ５には、中央部にその両側の２つの空間を仕切る１個の隔壁５１が形成されている。
ヘッダ４の２個の隔壁４１で仕切られる３個の空間のうち、両側の空間には、それぞれ冷媒流入口６、冷媒流出口７が接続される。
The header 4 is formed with two partition walls 41 that divide the central portion and the three spaces on both sides thereof, and the header 5 is formed with one partition wall 51 that divides the two spaces on both sides in the center portion. Yes.
Of the three spaces partitioned by the two partition walls 41 of the header 4, the refrigerant inlet 6 and the refrigerant outlet 7 are connected to the spaces on both sides, respectively.
冷媒は、図１において、冷媒流入口６からヘッダ４内の図示左側の空間４２に流入し、該空間４２に連通する複数の伝熱プレート２内の冷媒通路２３を下方に流動した後、ヘッダ５の隔壁５１より図示左側の空間５２に流入する。 In FIG. 1, the refrigerant flows into the space 42 on the left side of the header 4 in the header 4 from the refrigerant inlet 6 and flows downward through the refrigerant passages 23 in the plurality of heat transfer plates 2 communicating with the space 42. 5 flows into the space 52 on the left side of the figure.
次いで、図示右側に流動した後、隣接する複数の伝熱プレート２内の冷媒通路２３を上方に流動し、ヘッダ４内の図示中央の空間４３に流入する。
次いで、図示右側に流動した後、隣接する複数の伝熱プレート２内の冷媒通路２３を下方に流動した後、ヘッダ５の隔壁５１より図示右側の空間５３に空間に流入する。
Next, after flowing to the right side in the figure, it flows upward in the refrigerant passages 23 in the plurality of adjacent heat transfer plates 2 and flows into the center space 43 in the header 4 in the figure.
Next, after flowing to the right side in the figure, the refrigerant flows in the refrigerant passages 23 in the adjacent plurality of heat transfer plates 2 and then flows into the space 53 on the right side of the figure from the partition wall 51 of the header 5.
次いで、図示右側に流動した後、隣接する複数の伝熱プレート２内の冷媒通路２３を上方に流動し、ヘッダ４内の図示右側の空間４４に流入した後、冷媒流出管７から流出する。 Next, after flowing to the right side in the figure, the refrigerant flows in the refrigerant passages 23 in the plurality of adjacent heat transfer plates 2, flows into the space 44 on the right side in the figure in the header 4, and then flows out from the refrigerant outflow pipe 7.
一方、図２に示すように、図示しない送風ファンから送風される空気は、熱交換器１の伝熱プレート２の並び方向と直交する方向から送風される。送風空気は、伝熱プレート２間の隙間を通り抜け、その間に冷媒通路２３を流通する冷媒と送風空気との間で熱交換が行われ、蒸発器である熱交換器１の冷媒は、空気から熱を受けてガス状の冷媒に気化される。 On the other hand, as shown in FIG. 2, air blown from a blower fan (not shown) is blown from a direction orthogonal to the arrangement direction of the heat transfer plates 2 of the heat exchanger 1. The blown air passes through the gap between the heat transfer plates 2, and heat exchange is performed between the refrigerant flowing through the refrigerant passage 23 and the blown air, and the refrigerant of the heat exchanger 1, which is an evaporator, It receives heat and is vaporized into a gaseous refrigerant.
かかる構成を有した熱交換器１の各伝熱プレート２において、熱交換性能改善のため以下の考察を行った。
図４は、隣接する伝熱プレート２間に形成される空気流路の外郭線を抽出して示した図である。
In each heat transfer plate 2 of the heat exchanger 1 having such a configuration, the following consideration was made to improve the heat exchange performance.
FIG. 4 is a diagram showing an outline of an air flow path formed between adjacent heat transfer plates 2.
図４において、凸部の幅をｌ、凸部間ピッチをＬ、凸部の基準面からの高さをａｕ、凹部の基準面からの深さをａｌ、凸部２１及び凹部２２のテーパ角（テーパ面の基準面に対する垂線からの角度）をα、前記基準面と該基準面に対向する伝熱プレート２の基準面との距離をｄ１（＝２ｈ）とする。 In FIG. 4, the width of the convex portion is l, the pitch between the convex portions is L, the height of the convex portion from the reference plane is a u , the depth of the concave portion from the reference plane is a l , and the convex portion 21 and the concave portion 22 The taper angle (angle from the perpendicular to the reference surface of the taper surface) is α, and the distance between the reference surface and the reference surface of the heat transfer plate 2 facing the reference surface is d 1 (= 2h).
図４に示される凹凸を有した空気流路を流れる空気の熱伝達率、圧力損失の特性を求めるため、次式(1)，(2)のように定義される熱伝達率のｊファクター、及び圧力損失の指標となるｆファクターを、上記の各諸元ｌ、Ｌ、ａｕ、ａｌ、α、ｈの値を変更しつつＣＦＤ解析により求めた。 In order to obtain the heat transfer coefficient and pressure loss characteristics of the air flowing through the air flow path having unevenness shown in FIG. 4, the j factor of the heat transfer coefficient defined as the following equations (1) and (2): In addition, the f factor serving as an index of pressure loss was obtained by CFD analysis while changing the values of each of the above-mentioned specifications l, L, a u , a l , α, h.
数式(1)中でαは熱伝達率、ρｍは空気の入口〜出口の平均密度、Vは空気の流速、Cpは空気の定圧比熱、Prはプラントル数を示す。また、数式(2)中でρｉｎは空気の入口密度、ΔPは圧力損失、Kcは流入損失係数、Keは流出損失係数、σは前面面積と流体の通過面積の比を示す。 In the formula (1), α is a heat transfer coefficient, ρ m is an average density of air inlet to outlet, V is a flow velocity of air, Cp is a constant pressure specific heat of air, and Pr is a Prandtl number. In Equation (2), ρ in is the air inlet density, ΔP is the pressure loss, K c is the inflow loss coefficient, Ke is the outflow loss coefficient, and σ is the ratio of the front area to the fluid passage area.
ＣＦＤ解析で得られた多数のｊファクター及びｆファクターの値と、上記諸元で定まる凸部高さａｕ、凹部深さａｌ、テーパ角α等のパラメータとの間には、明確な相関関係は確認できなかった。すなわち、これらパラメータのみでは、ｊファクター，ｆファクターを精度よく示す相関式を得ることは難しく、熱交換器性能の向上をこれらパラメータだけの設定で予測することが難しいことが判明した。 There is a clear correlation between the values of many j-factors and f-factors obtained by CFD analysis and parameters such as the convex height a u , the concave depth a l and the taper angle α determined by the above specifications. The relationship could not be confirmed. That is, it has been found that with these parameters alone, it is difficult to obtain a correlation equation that accurately indicates the j-factor and f-factor, and it is difficult to predict the improvement of the heat exchanger performance by setting only these parameters.
そこで、ｊファクター，ｆファクターを精度よく示す相関式を得るため、凹凸を有した空気流路の熱流束分布の解析結果等に基づき、以下の３個のパラメータを新たに設定した。 Therefore, in order to obtain a correlation equation that accurately indicates the j factor and the f factor, the following three parameters were newly set based on the analysis result of the heat flux distribution of the air flow path having irregularities.
まず、伝熱プレート２の凸部２１の空気流通方向上流側の角部と、該伝熱プレート２に対向する伝熱プレート２の前記凸部の角部に近接する凸部２１の角部との間を流通する空気流の主軸が、前記対向する伝熱プレート２１相互を結ぶ垂直方向に対する鋭角側の「なす角θ」を第１のパラメータとして設定した。この「なす角」は、以下のように算出される。 First, the corner on the upstream side in the air flow direction of the convex portion 21 of the heat transfer plate 2 and the corner of the convex portion 21 adjacent to the corner of the convex portion of the heat transfer plate 2 facing the heat transfer plate 2 The first parameter is set as “the angle θ” formed by the main axis of the airflow flowing between the first and second heat transfer plates 21 at an acute angle with respect to the vertical direction connecting the opposing heat transfer plates 21 to each other. This “angle” is calculated as follows.
図４（Ａ）において、基準面間の流れは一様な流れであると仮定し、凸部頂面より上側の流れは基準面と平行に直進する流れになるとし、凸部頂面より下側の流れは凸部のテーパ角度に沿って流れるとする。隣接する伝熱プレート２間を流れる流量を１とすれば、上側流れベクトルｕ１、下側流れベクトルｕ２及びそれらの合成ベクトルｕ３は、それぞれ、下式(3)〜(5)で求められる。
したがって、式(5)を用いて、なす角θは、式(6)で算出される。
In FIG. 4 (A), it is assumed that the flow between the reference surfaces is a uniform flow, and the flow above the top surface of the convex portion is assumed to flow straight in parallel with the reference surface, and below the top surface of the convex portion. The flow on the side is assumed to flow along the taper angle of the convex portion. Assuming that the flow rate between adjacent heat transfer plates 2 is 1, the upper flow vector u 1 , the lower flow vector u 2 and their combined vector u 3 are obtained by the following equations (3) to (5), respectively. It is done.
Therefore, using equation (5), the angle θ formed is calculated by equation (6).
このなす角をパラメータ（変数）とすることで、空気が基準面間から凸部２１と凹部２２のテーパ面間に湾曲して流れる影響を評価することができると考えられる。
次に、図４（Ｂ）において、伝熱プレートの凸部の空気流通方向上流側の角部と、該伝熱プレートに対向する伝熱プレートの前記凸部の角部に近接する凸部の角部との距離ｄｃと、前記対向する各伝熱プレートの前記各角部の空気流通方向上流側に隣接して対向する面相互の距離ｄ１（＝２ｈ）とで定義される「縮流比β（＝ｄｃ／ｄ１）」を、第２のパラメータとして設定する。
By using this angle as a parameter (variable), it is considered that it is possible to evaluate the influence of the air flowing between the reference surface and the curved surface between the convex portion 21 and the concave portion 22.
Next, in FIG. 4B, the corners on the upstream side in the air flow direction of the convex portions of the heat transfer plate and the convex portions adjacent to the corners of the convex portions of the heat transfer plate facing the heat transfer plate The distance d c between the corners and the distance d 1 (= 2h) between the opposing faces adjacent to the upstream side in the air flow direction of the corners of the opposing heat transfer plates. The flow ratio β (= d c / d 1 ) ”is set as the second parameter.
前記角部同士の距離ｄｃ方向と対向する基準面同士の距離２ｈ方向とのなす角度をα’としたとき、この「縮流比β」は、以下のように算出される。 When the angle between the distance 2h direction of the reference surface facing each other and the distance d c direction between the corners and the alpha ', the' Chijimiryuhi β "is calculated as follows.
この縮流比βにより、基準面間を通過した流れが凹凸部開始点において縮流する影響を評価することができると考えられる。
次に、図４（Ｂ）において、距離ｄｃと、対向する各伝熱プレートの前記各角部の空気流通方向下流側に隣接して対向する面相互の距離ｄ２とで定義される「膨張比γ（＝ｄ２／ｄｃ）」を、第３のパラメータとして設定する。この「膨張比γ」は、以下のように算出される。
With this contraction ratio β, it is considered that the influence of the contraction of the flow that has passed between the reference planes at the starting point of the uneven portion can be evaluated.
Next, in FIG. 4 (B), it is defined by the distance d c and the distance d 2 between the surfaces facing each other adjacent to the downstream side in the air flow direction of each corner of each heat transfer plate. The expansion ratio γ (= d 2 / d c ) ”is set as the third parameter. This “expansion ratio γ” is calculated as follows.
この膨張比γにより、上記のように空気の流れが凹凸部開始点において縮流した後、凸部頂面と凹部底面との間で再度膨張することの影響を評価することができると考えられる。 With this expansion ratio γ, it is considered that the influence of expansion again between the top surface of the convex portion and the bottom surface of the concave portion can be evaluated after the air flow is contracted at the starting point of the concave and convex portion as described above. .
これらのパラメータ、なす角θ，縮流比β，膨張比γについて、それぞれ各値を変更しときの、ｊファクター及びｆファクターのＣＦＤ解析値と比較した。
なす角θについては、なす角が大きくなるほどｊファクター、ｆファクターともに小さくなる傾向が示された。
These parameters, the angle θ formed, the contraction ratio β, and the expansion ratio γ, were compared with the CFD analysis values of the j factor and f factor when each value was changed.
Regarding the formed angle θ, the j-factor and the f-factor tend to decrease as the formed angle increases.
縮流比βについては、縮流比βが小さくなるほどｊファクター、ｆファクターともに大きくなる傾向が示された。
膨張比γについては、膨張比γが大きくなるほどｊファクター、ｆファクターともに大きくなる傾向が示された。
As for the contraction ratio β, the j factor and the f factor tend to increase as the contraction ratio β decreases.
As for the expansion ratio γ, the j-factor and the f-factor tend to increase as the expansion ratio γ increases.
したがって、これらのパラメータを変数として用いることにより、ｊファクター及びｆファクターの精度の高い相関式の作成が可能になると考えられた。
そこで、上記のなす角θ，縮流比β，膨張比γに、その他の形状として凹凸部幅、凹凸ピッチ、凹凸数を加え、さらに流速に起因するＲｅ数（レイノルズ数）を用い、合わせて７個の変数を用いて、ｊファクター及びｆファクターの相関式を作成した。なお、なす角θ，縮流比β，膨張比γの中には、凸部高さａｕ、凹部深さａｌ、テーパ角αのパラメータが含まれる。
Therefore, it was considered that the use of these parameters as variables makes it possible to create a correlation equation with high accuracy of j-factor and f-factor.
Therefore, in addition to the angle θ, the contraction ratio β, and the expansion ratio γ, the width of the concavo-convex portion, the concavo-convex pitch, and the number of concavo-convex are added as other shapes, and the Re number (Reynolds number) resulting from the flow velocity is used together Using 7 variables, a correlation equation of j-factor and f-factor was created. It should be noted that the angle θ, the contraction ratio β, and the expansion ratio γ include parameters of the convex portion height a u , the concave portion depth a l , and the taper angle α.
具体的には、ｊファクター相関式、ｆファクター相関式を、Ｒｅ数の関数として(11),(12)式のように整理し、各式中の係数を形状パラメータの関数として式(13),(14)のように整理することとした。 Specifically, the j-factor correlation equation and the f-factor correlation equation are arranged as a function of Re number as shown in equations (11) and (12), and the coefficient in each equation is expressed as a function of shape parameter (13) , (14).
式(11),(12)中の各係数は、ＣＦＤ解析値との偏差が最も小さくなるようにし、最小二乗法を用いて数式(13),(14)中の定数Ｃ０〜Ｃ６を決定した。その結果、各係数の具体的な値は表１，２のようになる。また、表中の臨界Ｒｅ数は乱数への遷移を示す指標であり、数式(15)で推測される値を用いた。 Each coefficient in the equations (11) and (12) is set so that the deviation from the CFD analysis value is minimized, and the constants C 0 to C 6 in the equations (13) and (14) are calculated using the least square method. Were determined. As a result, specific values of each coefficient are as shown in Tables 1 and 2. Further, the critical Re number in the table is an index indicating a transition to a random number, and a value estimated by Equation (15) is used.
このように、最適な係数にフィッティングすることにより、図５、６に示すようにjファクター、fファクターともに、ＣＦＤ解析値との偏差を±15%以内に整理することができ、非常に予測精度の高い相関式を得ることができた。 In this way, by fitting to the optimal coefficient, the deviation from the CFD analysis value for both j factor and f factor can be arranged within ± 15% as shown in FIGS. A high correlation equation was obtained.
これにより、空気側の熱伝達率、圧力損失について、上記のｊファクター相関式とｆファクター相関式を用いて、高精度に推定できることが可能となった。しかし、空気側の熱伝達率、圧力損失だけを最適化したとしても、熱交換器全体の性能を最適化できるものではない。特に、本発明のようなフィンレス熱交換器においては、空気側の流路形状が冷媒側の流路形状に大きく影響を与えるため、冷媒側の流れも考慮した最適設計を実施する必要がある。 This makes it possible to estimate the air-side heat transfer coefficient and pressure loss with high accuracy using the above-described j-factor correlation equation and f-factor correlation equation. However, even if only the heat transfer coefficient and pressure loss on the air side are optimized, the performance of the entire heat exchanger cannot be optimized. In particular, in the finless heat exchanger as in the present invention, since the air-side flow channel shape greatly affects the refrigerant-side flow channel shape, it is necessary to implement an optimum design that also takes into account the refrigerant-side flow.
熱交換器全体の性能は、主に熱交換量、通気抵抗、冷媒側圧力損失の３つで評価される。熱交換量は大きく、通気抵抗、冷媒側圧力損失は小さくすることが望ましいが、これらは基本的に相反関係にあり、熱伝達率を高めて熱交換量を大きくすれば、圧力損失も大きくなる。したがって、熱交換器全体の性能を評価するためには、熱交換量、通気抵抗、冷媒側圧力損失の各性能を１つにまとめて評価する必要がある。 The overall performance of the heat exchanger is evaluated mainly by the heat exchange amount, ventilation resistance, and refrigerant side pressure loss. Although it is desirable that the heat exchange amount is large and the ventilation resistance and refrigerant side pressure loss are small, these are basically a reciprocal relationship, and if the heat exchange rate is increased to increase the heat exchange amount, the pressure loss also increases. . Therefore, in order to evaluate the performance of the entire heat exchanger, it is necessary to collectively evaluate each performance of heat exchange amount, ventilation resistance, and refrigerant side pressure loss.
まず、空気側の圧力損失（通気抵抗）の評価方法について一例を示す。送風ファンのＰＱ（差圧−風量）特性と通気抵抗との間には、図７に示すような関係がある。図中の実線はファンのＰＱ特性を示す。また、破線と一点鎖線は熱交換器Ａ，Ｂの通気抵抗を示し、これら２本の線の交点がそのファンを用いた場合の熱交換器に流れる動作点風量となる。熱交換器の通気抵抗が大きいと交点は低風量側に移動し、通気抵抗が小さいと高風量側に移動する。一方で、図８は風量と熱交換器の能力の関係を示している。つまり、熱交換器Ａは同一風量での能力は高いが、通気抵抗も高い熱交換器である。 First, an example of a method for evaluating air-side pressure loss (airflow resistance) will be shown. There is a relationship as shown in FIG. 7 between the PQ (differential pressure-air volume) characteristics of the blower fan and the ventilation resistance. The solid line in the figure shows the PQ characteristics of the fan. The broken line and the alternate long and short dash line indicate the ventilation resistance of the heat exchangers A and B, and the intersection of these two lines is the operating point airflow flowing through the heat exchanger when the fan is used. When the ventilation resistance of the heat exchanger is large, the intersection moves to the low air volume side, and when the ventilation resistance is small, the intersection moves to the high air volume side. On the other hand, FIG. 8 shows the relationship between the air volume and the capacity of the heat exchanger. That is, the heat exchanger A is a heat exchanger that has a high capacity with the same air volume but also has a high ventilation resistance.
ここで、図８中の動作点風量で能力を比較すると、熱交換器Ａは動作点での風量が小さくなるため、実際の能力は熱交換器Ｂよりも低くなる。このように、動作点風量での能力比較を行うことで、熱交換量と通気抵抗を１つにまとめて評価することができる。 Here, when the capacities are compared with the operating point air volume in FIG. 8, the actual capacity is lower than that of the heat exchanger B because the heat exchanger A has a small air volume at the operating point. In this way, by comparing the capacities with the air flow rate at the operating point, the heat exchange amount and the ventilation resistance can be collectively evaluated.
一方、冷媒側の圧力損失については、蒸発器を例にとると、蒸発器出口の圧力、温度条件を固定した評価を行うことで、圧力損失の影響が評価できる。蒸発器入口の圧力を基準とした場合、圧力損失の大きい熱交換器であれば、蒸発器入口の圧力は高くなり、熱交換器内を流れる冷媒の飽和温度も高くなる。熱交換量は熱伝達率と伝熱面積及び温度差の積で決まるため、このような条件では圧力損失の影響が熱交換能力に含まれていることになる。 On the other hand, with respect to the pressure loss on the refrigerant side, if an evaporator is taken as an example, the influence of the pressure loss can be evaluated by performing evaluation with the pressure and temperature conditions at the outlet of the evaporator being fixed. When the pressure at the evaporator inlet is used as a reference, if the heat exchanger has a large pressure loss, the pressure at the evaporator inlet increases and the saturation temperature of the refrigerant flowing in the heat exchanger also increases. Since the heat exchange amount is determined by the product of the heat transfer coefficient, the heat transfer area, and the temperature difference, the influence of pressure loss is included in the heat exchange capacity under such conditions.
以上のような方法であれば、通気抵抗や圧力損失を考慮した熱交換量の比較となり、熱交換器の最適化が可能になる。そこで、このような評価条件の下で、ｊファクターの相関式及びｆファクターの相関式を用いた熱交換器の性能予測シミュレーションによりパラメータスタディを実施した。その結果、得られた熱交換量を新たに設定したパラメータである、なす角θ、縮流比β、膨張比γを横軸にとって整理すると、図９〜図１１が求められる。同様に、凹凸ピッチ比ε、板厚ｔを横軸にとって整理すると図１２、図１３が求められる。 If it is the above methods, it will become the comparison of the heat exchange amount in consideration of ventilation resistance and pressure loss, and optimization of a heat exchanger will be attained. Therefore, under such evaluation conditions, a parameter study was performed by a performance prediction simulation of a heat exchanger using a j-factor correlation formula and an f-factor correlation formula. As a result, when the obtained angle of heat exchange, the formed angle θ, the contraction ratio β, and the expansion ratio γ are arranged on the horizontal axis, FIGS. 9 to 11 are obtained. Similarly, when the uneven pitch ratio ε and the plate thickness t are arranged on the horizontal axis, FIGS. 12 and 13 are obtained.
単位体積当たり熱交換量150[KW/m３]近傍を良好な熱交換性能を確保できる下限値として設定し、上記各パラメータ毎に、良好な熱交換性能が得られる数値範囲が求められた。
なす角については、図９に示すように、なす角θが０．２５〜０．６５[rad]であるときに単位体積当たり熱交換量が下限値以上となって良好な熱交換性能が得られることが明らかとなった。
The vicinity of the heat exchange amount per unit volume of 150 [KW / m 3 ] was set as a lower limit value for ensuring good heat exchange performance, and a numerical range for obtaining good heat exchange performance was obtained for each of the above parameters.
As for the formed angle, as shown in FIG. 9, when the formed angle θ is 0.25 to 0.65 [rad], the heat exchange amount per unit volume becomes equal to or higher than the lower limit value, and good heat exchange performance is obtained. It became clear that
縮流比βについては、図１０に示すように、縮流比βが０．１４〜０．４２であるときに単位体積当たり熱交換量が下限値以上となって良好な熱交換性能が得られることが明らかとなった。 Regarding the contraction ratio β, as shown in FIG. 10, when the contraction ratio β is 0.14 to 0.42, the heat exchange amount per unit volume becomes equal to or more than the lower limit value, and good heat exchange performance is obtained. It became clear that
膨張比γについては、図１１に示すように、膨張比γが２．００〜４．３０であるときに単位体積当たり熱交換量が下限値以上となって良好な熱交換性能が得られることが明らかとなった。 As for the expansion ratio γ, as shown in FIG. 11, when the expansion ratio γ is 2.00 to 4.30, the heat exchange amount per unit volume is equal to or higher than the lower limit value, and good heat exchange performance is obtained. Became clear.
ピッチ比εについては、図１２に示すように、ピッチ比εが４．０以下であるときに単位体積当たり熱交換量が下限値以上となって良好な熱交換性能が得られることが明らかとなった。 With respect to the pitch ratio ε, as shown in FIG. 12, it is clear that when the pitch ratio ε is 4.0 or less, the heat exchange amount per unit volume becomes equal to or greater than the lower limit value, and good heat exchange performance can be obtained. became.
板厚ｔについては、図１３に示すように、板厚ｔが０．３[mm]以下であるときに単位体積当たり熱交換量が下限値以上となって良好な熱交換性能が得られることが明らかとなった。 As for the plate thickness t, as shown in FIG. 13, when the plate thickness t is 0.3 [mm] or less, the heat exchange amount per unit volume is equal to or greater than the lower limit value, and good heat exchange performance is obtained. Became clear.
このように、なす角θ、縮流比β、膨張比γ、ピッチ比ε、板厚ｔをそれぞれ上記の数値範囲内とする伝熱プレート２の形状及び配置（プレート間隔）に設定した熱交換コアを形成することで、熱交換器の熱交換性能を最適化することができる。 In this way, the heat exchange is set to the shape and arrangement (plate interval) of the heat transfer plate 2 in which the angle θ, the contraction ratio β, the expansion ratio γ, the pitch ratio ε, and the plate thickness t are within the above numerical ranges, respectively. By forming the core, the heat exchange performance of the heat exchanger can be optimized.
なお、これらなす角θ、縮流比β、膨張比γ、ピッチ比ε、板厚ｔをそれぞれ上記の数値範囲内を満たしていれば、図３に示すように、伝熱プレート２の表面に形成される凸部と裏面に形成される凸部の形状は、異なっていてもよい。 If the angle θ, the contraction ratio β, the expansion ratio γ, the pitch ratio ε, and the plate thickness t are within the above numerical ranges, respectively, the surface of the heat transfer plate 2 as shown in FIG. The shape of the convex part formed in a back surface and the convex part formed in a back surface may differ.
１…熱交換器、２…伝熱プレート、３…熱交換用コア部、２１…凸部、２２…凹部、２３…冷媒通路、ｌ…凸部の幅、Ｌ…凸部のピッチ、ａｕ…凸部の高さ、ａｌ…凹部の深さ、ｄ１…伝熱プレートの基準面間距離（＝２ｈ）、α…凸部及び凹部のテーパ角、θ…なす角、β…縮流比、γ…膨張比、ピッチ比ε、板厚ｔ DESCRIPTION OF SYMBOLS 1 ... Heat exchanger, 2 ... Heat transfer plate, 3 ... Heat exchange core part, 21 ... Convex part, 22 ... Concave part, 23 ... Refrigerant passage, l ... Width of convex part, L ... Pitch of convex part, au ... height of convex part, a 1 ... depth of concave part, d 1 ... distance between reference surfaces of heat transfer plate (= 2h), α ... taper angle of convex part and concave part, θ ... angle formed, β ... contraction flow Ratio, γ ... expansion ratio, pitch ratio ε, plate thickness t
Claims (4)
前記各伝熱プレートの凸部を、対向する伝熱プレートの凸部間の凹部に対向するように配設し、
前記各伝熱プレート相互間に、前記一方向と平行な方向に流体を流通させるように積層して構成した熱交換用コア部を備えた熱交換器において、
前記伝熱プレートの凸部の流体流通方向上流側の角部と、該伝熱プレートに対向する伝熱プレートの前記凸部の角部に近接する凸部の角部との距離ｄｃと、前記対向する各伝熱プレートの前記各角部の流体流通方向上流側に隣接して対向する面相互の距離ｄ１とで定義される縮流比β（＝ｄｃ／ｄ１）が、０．１４〜０．４２の範囲内の値に設定され、
前記縮流比βは、前記伝熱プレートの凸部の基準面からの高さをａｕ、前記基準面と該基準面に対向する伝熱プレートの基準面との距離を２ｈ、前記角部同士の距離ｄｃ方向と前記対向する基準面同士の距離２ｈ方向とのなす角度をα’としたとき、下記の式で算出されることを特徴とする熱交換器。
The convex portions of the heat transfer plates are disposed so as to face the concave portions between the convex portions of the opposing heat transfer plates,
Between the heat transfer plates, a heat exchanger including a heat exchanging core portion configured to be laminated so that a fluid flows in a direction parallel to the one direction,
And the heat transfer corner of the fluid flow direction upstream side of the convex portion of the plate, and the distance d c between the convex corner portion adjacent to the corner portion of the convex portion of the heat transfer plate facing the heat transfer plate, The contraction ratio β (= d c / d 1 ) defined by the distance d 1 between the opposing surfaces adjacent to the upstream side in the fluid flow direction of the corners of the opposing heat transfer plates is 0. Set to a value in the range of .14 to 0.42,
The contraction ratio β is the height of the convex portion of the heat transfer plate from the reference surface a u , the distance between the reference surface and the reference surface of the heat transfer plate facing the reference surface is 2 h, and the corner portion when the angle between the distance 2h direction of the reference faces of the distance d c direction between the opposing was alpha ', the heat exchanger, characterized in that it is calculated by the following equation.
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