JP2005048608A

JP2005048608A - Splitter runner and hydraulic machinery

Info

Publication number: JP2005048608A
Application number: JP2003203613A
Authority: JP
Inventors: Hideyuki Kawajiri; 秀之川尻; Kiyoshi Matsumoto; 貴與志松本; Takanori Nakamura; 高紀中村; Kotaro Tezuka; 光太郎手塚; Takeo Tokumiya; 健男徳宮
Original assignee: Toshiba Corp
Current assignee: Toshiba Corp
Priority date: 2003-07-30
Filing date: 2003-07-30
Publication date: 2005-02-24
Anticipated expiration: 2023-07-30
Also published as: JP4163062B2

Abstract

<P>PROBLEM TO BE SOLVED: To provide a splitter runner and a hydraulic machinery improving cavitation performance at the time of a head difference operation. <P>SOLUTION: At least a part of the water wheel inlet side end of a short blade is positioned on the inner diameter side of a periphery of the water wheel inlet side end of a long blade. The blade loads of the long and short blades are uniformalized by applying load to the short blade by a flow led to flow into the runner and letting a flow with pressure reduced to flow into the long blade. As a result, difference in deflecting force in both the blades is reduced, difference in an angle of incidence is reduced, and a friction loss is reduced in comparison with the same size runner by shortening the long blade, therefore, cavitation performance is improved. <P>COPYRIGHT: (C)2005,JPO&NCIPI

Description

【０００１】
【発明の属する技術分野】
本発明は、長翼の間に短翼を配したスプリッタランナおよび水力機械に関する。
【０００２】
【従来の技術】
水力発電（例えば、揚水発電）に、水の速度エネルギーおよび圧力エネルギーの双方を機械的エネルギーに変換する反動水車が用いられる。反動水車のうちフランシス形水車は、回転軸を中心とする円周方向に複数枚の羽根を配列したランナを有し、水が外周方向からランナへと流入することで、ランナが回転して水車出力が生成される。
フランシス形水車では、ランナの羽根を同一形状とするのが一般的であるが、水車効率、キャビテーション性能、低振動、変落差特性等の水車性能の向上を図るため、羽根を交互に長翼、短翼とするスプリッタランナを用いる場合がある。スプリッタランナでは、多翼化による整流効果により、翼間での２次流れが低減され、水力効率を向上できる。また、羽根一枚当たりの翼負荷が低減されることで、キャビテーション性能を向上できる。なお、スプリッタランナに関する技術が公開されている（特許文献１，２参照）。
【０００３】
【特許文献１】
特開２０００−５４９４４
【特許文献２】
特開２０００−２０５１０１
【０００４】
【発明が解決しようとする課題】
しかしながら、従来のスプリッタランナでは、羽根の特性（例えば、キャビテーション特性）が、長翼と短翼とで異なっている。
例えば、設計点に比べて水の落差が大きい高落差運転時には長翼でキャビテーションが生じ易くなり、設計点に比べて水の落差が小さい低落差運転時には短翼でキャビテーションが生じ易くなる。このような場合には、落差が変動する変落差運転時において、キャビテーションを生じずに運転可能な範囲が制限される。
また、長翼と短翼の特性の相違が、低落差領域や高落差領域での効率低下の原因となる可能性もある。
上記に鑑み、本発明は変落差運転時におけるキャビテーション性能の向上を図ったスプリッタランナおよび水力機械を提供することを目的とする。
【０００５】
【課題を解決するための手段】
（１）上記目的を達成するために、本発明に係るスプリッタランナは、クラウン、バンド間に、翼長の長い羽根と翼長の短い羽根とが、回転軸を中心とする円周方向に交互に配置されてなるスプリッタランナであって、前記翼長の長い羽根の前記クラウン、バンド間の水車入口側端部の少なくとも一部が、前記翼長の短い羽根の前記クラウン、バンド間の水車入口側端部よりも、前記円周の内径側に位置することを特徴とする。
【０００６】
ランナに流入してくる流れが外径側に位置する翼長の短い羽根（短翼）に負荷を与え、圧力が減少した流れが翼長の長い羽根（長翼）に流入する。このため、短翼と長翼双方の水車入口側端部が同一円周上に位置する場合に比べて、短翼が受け持つ負荷が増加し、長翼が受け持つ負荷が減少する。その結果、長翼と短翼の翼負荷が均一化されて、両翼の水車入口側端部に作用する偏向力の差異が低減し、迎え角度の相違が小さくなる。その結果、キャビテーション特性を向上できる。
また、長翼を短くすることにより、同サイズのランナと比べて摩擦損失が低減され、水車効率の向上が図られる。
【０００７】
（２）本発明に係るスプリッタランナは、クラウン、バンド間に、翼長の長い羽根と翼長の短い羽根とが、回転軸を中心とする円周方向に交互に配置されてなるスプリッタランナであって、前記翼長の長い羽根の厚みの流路上での最大値Ｔａが、前記翼長の短い羽根の厚みの流路上での最大値Ｔｂよりも小さいことを特徴とする。
【０００８】
翼長の短い羽根（短翼）の厚みの最大値Ｔａが、翼長の長い羽根（長翼）の厚みの最大値Ｔｂより大きいことから、短翼まわりの循環が長翼まわりの循環に近くなり、短翼と長翼の翼負荷の差異を小さくすることができる。このため、両翼の水車入口端で作用する偏向力の差異が低減し、迎え角度の相違が小さくなり、キャビテーション特性を向上できる。
【０００９】
（３）本発明に係るスプリッタランナは、クラウン、バンド間に、翼長の長い羽根と翼長の短い羽根とが、回転軸を中心とする円周方向に交互に配置されてなるスプリッタランナであって、前記円周の外径方向からみて、前記翼長の長い羽根の水車入口側端部に沿う水車入口端曲線と前記バンドの面とがなす圧力面側の角度θａが、前記翼長の短い羽根の水車入口端曲線と前記バンドの面とがなす圧力面側の角度θｂよりも小さいことを特徴とする。
【００１０】
一般に、高落差運転時において、翼長の長い羽根（長翼）の負圧側水車入口端近傍は翼長の短い羽根（短翼）の負圧側水車入口端近傍よりもキャビテーションが発生し易い。
長翼の水車入口端曲線の角度θａを、短翼の水車入口端曲線の角度θｂよりも小さくすることで、長翼の負圧側水車入口端近傍での圧力低下を短翼に比べて緩和できる。このため、長翼の高落差側キャビテーション限界を、より限界値の大きい短翼の高落差側キャビテーション限界に近づけ、高落差運転時でのキャビテーション限界を広げることができる。
【００１１】
（４）本発明に係るスプリッタランナは、クラウン、バンド間に、翼長の長い羽根と翼長の短い羽根とが、回転軸を中心とする円周方向に交互に配置されてなるスプリッタランナであって、前記翼長の長い羽根の中心線と前記円周とが水車入口側においてなす角度βａの前記円周の半径ｒに対する変化率（∂βａ／∂ｒ）が、前記翼長の短い羽根の中心線と前記円周とが前記水車入口側においてなす角度βｂの前記円周の半径ｒに対する変化率（∂βｂ／∂ｒ）より大きいことを特徴とする。
【００１２】
翼長の長い羽根（長翼）周りの循環と翼長の短い羽根（短翼）周りの循環の違いにより、それぞれの羽根への局所的な流入角度の差が生じる。水車入口側付近の羽根の中心線を調整することで、流入流れに対する両翼の迎え角度の相違が小さくなり、キャビテーション特性を向上できる。
【００１３】
（５）本発明に係るスプリッタランナは、クラウン、バンド間に、翼長の長い羽根と翼長の短い羽根とが、回転軸を中心とする円周方向に交互に配置されてなるスプリッタランナであって、水車入口側において前記翼長の長い羽根の中心線と前記円周とがなす負圧面側の角度βａと、前記水車入口側において前記翼長の短い羽根の中心線と前記円周とがなす負圧面側の角度βｂの差（βａ−βｂ）が、２°以上、６°以下であることを特徴とする。
【００１４】
翼長の長い羽根（長翼）と翼長の短い羽根（短翼）の局所的な流入角度の相違に対応して羽根の水車入口端角度を調整することで、両翼の迎え角度の相違が小さくなり、キャビテーション特性を向上できる。
【００１５】
【発明の実施の形態】
以下、図面を参照して、本発明の実施の形態を詳細に説明する。
（第１の実施の形態）
図１は本発明の第１実施形態に係る水力機械１０を側面から見た状態を表す一部断面図である。また、図２は、水力機械１０のランナ４０を下方（吸出管５０側）から見た状態を表す一部断面図である。
水力機械１０は、ケーシング２０，ガイドベーン（案内羽根）３０，ランナ４０，吸出管５０，回転軸６０，発電電動機７０を備え、発電（揚水発電を含む）に用いることができる。即ち、ケーシング２０から水が流入し、ガイドベーン３０およびランナ４０を通って、吸出管５０に排出されることで、ランナ４０が回転されて発電が行われる。
水の落差（揚水発電では上池と下池の水位の相違）によりランナ４０が回転されるのであり、水力機械１０はいわゆるフランシス形水車として機能する。
【００１６】
ケーシング２０は、断面積がしだいに減少するドーナツ形状であり、発電時に上池から水が流入する。
ガイドベーン３０は、ケーシング２０からランナ４０に流入する水の流量を調節するものであり、ランナ４０の外側に円周方向に所定の間隔をおいて配置されている。ガイドベーン３０はそれぞれ、その中心のまわりに回転でき，それにより水の流量を調節できる。ガイドベーン３０によって、水流が円周方向の速度成分をもつ旋回流となり，外周方向からランナ４０に流入する。
【００１７】
ランナ４０は、クラウン４１、バンド４２，長翼４３，短翼４４を備え、ケーシング２０から流入する水によって回転される。長翼４３（短翼４４より長い翼長の羽根），短翼４４（長翼４３より短い翼長の羽根）は、クラウン４１、バンド４２間の円周方向に交互に配置される。即ち、ランナ４０はスプリッタランナである。なお、ランナ４０（特に、長翼４３，短翼４４）の詳細な形状は後述する。
【００１８】
吸出管５０は、ランナ４０の回転に用いられた水を放水路（揚水発電の場合は下池）へと流出させる。
回転軸６０は、ランナ４０の回転運動を発電電動機７０に伝達する。
発電電動機７０は、回転軸６０が回転されることで発電を行う。
【００１９】
（ランナ４０の形状の詳細）
図３（Ａ），（Ｂ）はそれぞれ、ランナ４０の子午面形状、および断面形状を表す図である。図３（Ａ）は、図１に対応し、ランナ４０を子午面（回転軸６０を含む面）から見た形状を表す。図３（Ｂ）は、長翼４３、短翼４４を図３（Ａ）の面Ｆ（回転軸６０に垂直な面）で切断した状態を表す。
図１〜３に示すように、短翼４４の水車入口端曲線Ｌｂに対し、長翼４３の水車入口端曲線Ｌａの少なくとも一部がランナ４０の内径側に位置している。
【００２０】
ここで、水車入口端曲線Ｌａ、Ｌｂはそれぞれ、長翼４３、短翼４４について、クラウン４１、バンド４２間の水車入口側（ケーシング２０側）の翼端に沿った曲線である。また、水車出口端曲線Ｅａ，Ｅｂはそれぞれ、長翼４３、短翼４４について、クラウン４１、バンド４２間の水車出口側（吸出管５０側）の翼端に沿った曲線である。円周Ｄａ１，Ｄｂ１はそれぞれ、長翼４３、短翼４４の水車入口側の複数の翼端を結ぶ円弧であり、回転軸６０をその中心とする。
【００２１】
（本実施形態における基本的な考え方）
本実施形態における基本的な考え方を説明する。
図４（Ａ），（Ｂ）はそれぞれ、本実施形態の比較例たるランナ４０ｘの子午面形状、および断面形状を表す図であり、図３（Ａ），（Ｂ）に対応する。
ランナ４０ｘでは、長翼４３ｘと短翼４４ｘは、翼長が互いに異なるものの、その翼形状が近似しており、回転軸６０中心から水車入口側の翼端までの距離が等しい。短翼４４ｘは、長翼４３ｘと別個に作成されるのではなく、長翼４３ｘを水車入口側から６５〜８０％程度の長さの部分で切り取ることで、作成されるのが通例だからである。
【００２２】
図５は、ランナ４０ｘに流入する水の流れを表す図である。
ランナ４０ｘに、上流（ケーシング２０側）から流入角度βｗで水が流入する。羽根（長翼４３ｘおよび短翼４４ｘ）の近傍では流入角度がΔβｗ変化して（βｗ＋Δβｗ）となる。これは、それぞれの羽根（長翼４３ｘおよび短翼４４ｘ）の周りの循環に基づく偏向力によるものである。
【００２３】
長翼４３と短翼４４では、断面形状が異なることから、翼負荷が異なる。一般に、長翼４３の翼負荷ΓＡが短翼４４の翼負荷ΓＢに比べて大きいことから、長翼４３近傍での流入角度の変化ΔβｗＡは短翼４４近傍での流入角度の変化ΔβｗＢより大きい（ΔβｗＡ＞ΔβｗＢ）。即ち、長翼４３への流入角度（βｗ＋ΔβｗＡ）は、短翼４４への流入角度（βｗ＋ΔβｗＢ）より大きくなる（βｗ＋ΔβｗＡ＞ βｗ＋ΔβｗＢ）。
【００２４】
長翼４３と短翼４４とで水の流入角度が異なることに起因して、ランナ４０ｘのキャビテーション特性の劣化が生じることが判った。以下に、長翼４３と短翼４４での流入角度の相違とランナ４０ｘのキャビテーション特性の関係につき説明する。
【００２５】
長翼４３と短翼４４での流入角度の相違は、迎え角（流入角度と羽根角度の差、正確には翼弦（水車入口側端部と水車出口側端部とを結ぶ羽根の中心線）が水の流れ方向（流線）となす角度）の相違をもたらす。
迎え角が大きいと羽根の水車入口側端部付近の翼面で静圧が低下し易くなる。従い、キャビテーション（静圧が局部的に蒸気圧以下になり，その部分の水が蒸発して水蒸気の気泡が生ずる現象）が生じやすくなる。
この結果、長翼４３と短翼４４とで迎え角が異なると、キャビテーションが発生する条件が長翼４３と短翼４４とで異なり、ランナ４０ｘ全体としてキャビテーションが発生し易くなる。長翼４３と短翼４４いずれでキャビテーションが発生しても、ランナ４０ｘにキャビテーションが生じることに変わりはないからである。
【００２６】
以上のように、長翼４３と短翼４４で水の流入角度（究極的には迎え角）が相違することで、ランナ４０ｘにキャビテーションが発生し易くなる。逆にいえば、長翼４３と短翼４４での迎え角の差異を小さくすることで、水車のキャビテーション性能等を向上できる。
本実施形態では、長翼４３の水車入口端曲線Ｌａの少なくとも一部を短翼４４の水車入口端曲線Ｌｂよりもランナ４０の内径側に位置させることで、長翼４３と短翼４４での迎え角の差異の解消を図っている。
【００２７】
（ランナ４０の特性）
以下、本実施形態に係るランナ４０の特性を比較例のランナ４０ｘと対比して説明する。
図６（Ａ）、（Ｂ）はそれぞれ、図４に示した比較例のランナ４０ｘ、図１〜３に示した本実施形態のランナ４０でのキャビテーション特性を表すグラフである。グラフの横軸が水の落差（上池と下池の水位の差）であり、縦軸がランナに流入する水の流量を表す。
破線が水車効率の等しい点を結んだ等効率曲線を表す。また、実線が長翼４３の水車入口（ケーシング２０側の翼端）でのキャビテーションの初生点を、一点鎖線が短翼４４の水車入口でのキャビテーションの初生点を表す。なお、「×」は、水車効率が最高になる最高効率点を表す。
【００２８】
図７（Ａ）、（Ｂ）はそれぞれ、図４に示した比較例のランナ４０ｘ、図１〜３に示した本実施形態のランナ４０の最高効率点付近での静圧分布を表すグラフである。また、図８（Ａ）、（Ｂ）はそれぞれ、高落差側、低落差側での比較例のランナ４０ｘの静圧分布を表すグラフである。図９（Ａ）、（Ｂ）はそれぞれ、高落差側、低落差側での本実施形態のランナ４０の静圧分布を表すグラフである。
図７〜９のグラフの横軸は翼面上の流線に沿った長さであり、縦軸は翼面の静圧を表す。
【００２９】
図６（Ａ）に示すように、比較例においては、高落差側でのキャビテーション初生点は長翼４３ｘの方が短翼４４ｘより低落差側である。逆に、低落差側でのキャビテーション初生点は短翼４４ｘの方が長翼４３ｘより高落差側である。このため、水車入口のキャビテーション無発生区間（最高効率点の流量でキャビテーションが発生しない落差の幅）Ｓ０が狭くなっている。
これに対して、図６（Ｂ）に示すように、本実施形態では長翼４３と短翼４４のキャビテーション初生点が高落差側、低落差側の双方で一致している。このため、水車入口のキャビテーション無発生区間（最高効率点の流量でキャビテーションが発生しない落差の幅）Ｓ１が広くなり、水車入口でのキャビテーション特性が向上する。
【００３０】
これは、図８，９に示すように、水車入口側の翼面の静圧の最低点が、比較例では長翼４３ｘ，短翼４４ｘで一致せず、本実施形態では長翼４３，短翼４４で一致することによる。図８から、高落差側では長翼４３ｘの水車入口側翼端で、低落差側では短翼４４ｘの水車入口側翼端で、それぞれの静圧が水の飽和蒸気圧より低下し、キャビテーションが発生することが判る。なお、図８、９では高落差、低落差の場合を示したがその中間においても長翼４３，短翼４４の静圧の一致、不一致の傾向には特に変わりがない。
なお、図６から、長翼４３と短翼４４の無衝突流入点が比較例では一致せず、本実施形態では一致していることも判る。
【００３１】
本実施形態に係るランナ４０では、上流（ケーシング２０側）から自由渦的に流入してくるほぼ圧力一定な流れが、ランナ４０の外径側に位置する短翼４４に作用することにより、短翼４４がまず負荷を持つ。従い、長翼４３には圧力が減少した流れが作用する。
このために、本実施形態に係るランナ４０では、短翼４４と長翼４３の入口端が同一円周上に位置する比較例に比べて、短翼４４が受け持つ負荷が増加し、逆に長翼４３が受け持つ負荷が減少する。従い、比較例では図７（Ａ）に示すように長翼４３ｘが大きな負荷を持ち、短翼４４ｘはあまり大きな負荷を持たないのに対して、本実施形態では図７（Ｂ）に示すように長翼４３と短翼４４の翼負荷が均一化される（長翼４３の静圧の積分値が、短翼４４の静圧の積分値に近づく）。
【００３２】
以上から、本実施形態では長翼４３、短翼４４の入口側翼端で作用する偏向力の差異が低減し、流れの流入角度（βｗ＋Δβｗ）の相違が小さくなる。従い、変落差運転時の入口側翼端における圧力低下レベルの差も小さくなる。
その結果、図６（Ｂ）に示すように、水車入口キャビテーションの初生点を長翼４３と短翼４４でほぼ同じ落差とすることができ、水車入口のキャビテーション性能の向上が可能となる。
これに加えて、長翼４３が短くなることにより、同サイズのランナ４０ｘと比べて、摩擦損失が減少することから、水力効率が向上する。
さらに、長翼４３と短翼４４の特性のずれが改善されて無衝突流入点が一致するため最高効率が向上する。
【００３３】
続いて、長翼４３と短翼４４の外径を変化させた場合を定量的に説明する。
図１０は、長翼４３と短翼４４の外径の比とキャビテーション無発生区間Ｓの関係を表すグラフである。グラフの横軸が、回転軸に垂直な断面における長翼４３の外径Ｄａ１と短翼４４の外径Ｄｂ１の比Ｄａ１／Ｄｂ１を表す。横軸が、長翼４３と短翼４４の外径が等しいときのキャビテーション無発生区間Ｓ０に対するＳとキャビテーション無発生区間Ｓ１の比Ｓ１／Ｓ０を表す。
【００３４】
長翼４３と短翼４４の外径の比（Ｄａ１／Ｄｂ１）が小さくなると、短翼４４周りの循環ΓＢが増し、長翼４３周りの循環ΓＡが減少する。このため、それぞれの羽根への局所的な流入角度の差が小さくなり、Ｓ１／Ｓ０が大きくなる。
一方、Ｄａ１／Ｄｂ１が小さくなりすぎると、逆に短翼４４が過大な負荷を持ち、短翼４４周りの循環ΓＢが長翼４３周りの循環ΓＡを上回ることとなる。このため、短翼４４と長翼４３の立場が逆転してＳ１／Ｓ０が減少に転じる。
そこで図１０に示すように、長翼４３の外径と短翼４４の外径の比Ｄａ１／Ｄｂ１を０．８５以上、０．９８以下の範囲とすることで（０．８５≦Ｄａ１／Ｄｂ１≦０．９８）、キャビテーション無発生区間Ｓを大きくすることが可能となる。
【００３５】
（第２の実施の形態）
本発明の第２の実施形態を説明する。
図１１は、本発明の第２の実施形態に係るランナ４０ａの断面形状を表す断面図である。本図は、図３（Ｂ）と異なり、水車入口（ケーシング２０側）から水車出口（吸出管５０側）への水の流路に沿う面で長翼４３と短翼４４とを切断し、これを展開した状態を表す。
【００３６】
本実施形態では、長翼４３の羽根最大厚みＴａが、短翼４４の羽根最大厚みＴｂより小さくなっている（Ｔａ＜Ｔｂ）。羽根最大厚みＴａ、Ｔｂは、羽根（長翼４３、短翼４４）を流路に沿う面で切断したときの羽根の厚みの最大値をいう。
羽根の厚みおよび翼長以外の条件は、長翼４３と短翼４４とで大きく変わることはない。即ち、本実施形態では、長翼４３と短翼４４の外径Ｄａ１、Ｄｂ１は等しいとしている。
これ以外の条件は第１の実施形態と本質的に異なる訳ではないので、全体の構成の図示、説明は省略する。
【００３７】
本実施形態において、長翼４３と短翼４４での迎え角の相違を低減することで、キャビテーション特性の向上を図っていることは、第１の実施形態と同様である。但し、本実施形態では、長翼４３と短翼４４の厚みに着目して長翼４３と短翼４４の迎え角の相違を低減している。
短翼４４の厚みを増すことで、短翼４４周りの循環が増加し、短翼４４がより大きな負荷を持つ。このため、短翼４４と長翼４３の翼負荷の差異が小さくなる（翼負荷の差異の均一化）。
この結果、長翼４３と短翼４４の水車入口側翼端に作用する偏向力の差異が低減し、迎え角度の相違が小さくなり、キャビテーション性能の向上が可能となる。
【００３８】
次に、長翼４３、短翼４４の羽根最大厚みＴａ、Ｔｂと水力損失の関係につき説明する。
図１２は、長翼４３に対する短翼４４の羽根最大厚みの比Ｔｂ／Ｔａと水力損失の関係の１例を表すグラフである。横軸が長翼４３の羽根最大厚みＴａに対する短翼４４の羽根最大厚みＴｂの比（Ｔｂ／Ｔａ）を、縦軸が水力損失を表す。なお、本図では流入流れと長翼４３とのマッチングがとれている運転点を例にとっている。
【００３９】
短翼４４周りの循環が増加することで、流入流れの角度変化Δβｗが大きくなり、短翼４４の無衝突流入角度付近で短翼４４の衝突損失が最小になる。短翼４４をさらに厚くして短翼４４回りの循環が増すと、短翼４４の衝突損失は増加に転じる。
一方、短翼４４が厚くなると、流路の断面積が減少し、平均流速が増加するために摩擦損失は増加する。
この結果、衝突損失と摩擦損失を合計した水力損失は、羽根最大厚みの比Ｔｂ／Ｔａが変化することで、減少し、その後に増加することとなる。
【００４０】
図１２に示すように、長翼４３と短翼４４の羽根最大厚みの比Ｔｂ／Ｔａを、１．０５以上、１．３以下（１．０５≦Ｔｂ／Ｔａ≦１．３）とすることで、合計の水力損失を最小限の範囲とすることができた。即ち、短翼４４をある程度厚くすることで、短翼４４の水力損失の低減を図ることができる。
この場合、短翼４４が厚くなることで、水力損失の低減と併せて、短翼４４の強度の向上をも図ることができる。
【００４１】
（第３の実施の形態）
本発明の第３の実施形態を説明する。
図１３は本発明の第３の実施形態に係るランナ４０ｂを側面（外径方向）から見た状態を表す側面図である。
クラウン４１とバンド４２の間に、長翼４３と、短翼４４が周方向に交互に配置されている。本実施形態では、長翼４３の水車入口端曲線Ｌａとバンド４２の面とがなす角θａより、短翼４４の水車入口端曲線Ｌｂとバンド４２の面とがなす角θｂが大きい。
【００４２】
正確には、角度θａは接線６ｂと接線６ｃとのなす角として、角度θｂは接線７ｂと接線７ｃとのなす角として定義できる。
接線６ｂは、外径方向から見た長翼４３の水車入口端曲線Ｌａとバンド４２との交点６ａにおける、水車入口端曲線Ｌａの接線である。接線６ｃは、交点６ａにおける、交点６ａを通り回転軸６０を中心とする円の回転方向と反対向きの接線である。
接線７ｂは、外径方向から見た短翼４４の水車入口端曲線Ｌｂとバンド４２との交点７ａにおける、水車入口端曲線Ｌｂの接線である。接線７ｃは、交点７ａにおける、交点７ａを通り回転軸６０を中心とする円の回転方向と反対向きの接線である。
【００４３】
水車入口端曲線の角度および翼長以外の条件は、長翼４３と短翼４４とで大きく変わることはない。即ち、本実施形態では、長翼４３と短翼４４の外径Ｄａ１、Ｄｂ１は等しいとしている。
これ以外の条件は第１の実施形態と本質的に異なる訳ではないので、全体の構成の図示、説明は省略する。
本実施形態においては、長翼４３と短翼４４の水車入口端曲線の角度に着目して長翼４３と短翼４４の迎え角の相違を低減している。
【００４４】
図１４は、長翼４３の水車入口端曲線の角度θａとキャビテーションの関係の１例を表すグラフである。グラフの横軸は長翼４３の水車入口端曲線の角度θａを表す。グラフの縦軸は長翼４３の負圧面での最低静圧Ｐｓａ−ｍｉｎ（破線のグラフ）および水力効率（実線のグラフ）を表す。ここでは、水車入口端近傍の負圧面でキャビテーションが発生し易い運転点を例に挙げている。
【００４５】
図１４から、長翼４３の水車入口端曲線の角度θａより短翼４４の水車入口端曲線の角度θｂを大きくすることで、水力効率をさほど低下させずに、キャビテーションを防止できることが判る（静水圧を飽和蒸気圧より大きくする）。
これは、長翼４３での角度θａを短翼４４での角度θｂより小さくすることで、長翼４３の水車入口端付近バンド４２側（交点６ａ付近）の負圧面での流れの偏りが緩和されることによる。流れの偏りの緩和によって、交点６ａ付近での局所的な圧力低下が抑制される。
【００４６】
短翼４４と長翼４３が水車入口側で同一の羽根形状を持つ場合には、図６（Ａ）および図８（Ａ）で示したように、高落差運転時において、短翼４４よりも低い落差で、長翼４３の水車入口側負圧面にキャビテーションが発生する。
これに対して、長翼４３の角度θａを短翼４４の角度θｂより小さくすることで、長翼４３での圧力低下を短翼４４に比較して緩和し、長翼４３のキャビテーション初生点をより高落差側に移すことができる。即ち、水車入口側負圧面でキャビテーションが発生する落差を短翼４４と長翼４３で一致させることができ、高落差側での運転範囲を拡大できる。
【００４７】
短翼４４、長翼４３の角度θｂ、θａは、水車入口での高落差側キャビテーション特性が仕様を満足するように設定される。短翼４４の角度θｂは、短翼４４の入口負圧面でキャビテーションが発生しないように定められる。そして、短翼４４、長翼４３の角度差（θｂ−θａ）をある程度大きくすることで、長翼４３の入口負圧面でのキャビテーションの発生を防止できる。
この一方、図１４に示すように、角度差（θｂ−θａ）を大きくし過ぎると水力効率が低下する。
図１４に示すように、角度差（θｂ−θａ）を５°以上３０°以下とすることで、飽和蒸気圧（キャビテーションが発生する圧力）に対して、静水圧のマージンを確保し、かつ水力効率を損なわないことが可能となる。
【００４８】
（第４の実施の形態）
本発明の第４の実施形態を説明する。
図１５は、本発明の第４の実施の形態に係るランナ４０ｃの断面形状を表す断面図である。本図は、水車入口（ケーシング２０側）から水車出口（吸出管５０側）への水の流路に沿う面で長翼４３と短翼４４とを切断し、これを展開した状態を表す。
【００４９】
本実施形態では、水車入口側において、長翼４３の反り量Ｃａが短翼４４の反り量Ｃｂより大きい。
ここで、長翼４３の反り量Ｃａは、長翼４３のキャンバーライン４５（流路に沿って、水車入口側端部と水車出口側端部とを結ぶ長翼４３の中心線）と回転軸６０を中心とする半径ｒの円とのなす角度βａの半径ｒに対する変化率（∂βａ／∂ｒ）として定義できる。
また、短翼４４の反り量Ｃｂは、短翼４４のキャンバーライン４６（流路に沿って、水車入口側端部と水車出口側端部とを結ぶ短翼４４の中心線）と半径ｒの円とのなす角度βｂの半径ｒに対する変化率（∂βｂ／∂ｒ）として定義できる。
【００５０】
羽根の反り量および翼長以外の条件は、長翼４３と短翼４４とで大きく変わることはない。即ち、本実施形態では、長翼４３と短翼４４の外径Ｄａ１、Ｄｂ１は等しいとしている。
これ以外の条件は第１の実施形態と本質的に異なる訳ではないので、全体の構成の図示、説明は省略する。
【００５１】
本実施形態において、長翼４３と短翼４４での迎え角の相違を低減することで、キャビテーション特性の向上を図っていることは、第１の実施形態と同様である。但し、本実施形態では、長翼４３と短翼４４の反り量に着目して長翼４３と短翼４４の迎え角の相違を低減している。
図５で示したように、長翼４３と短翼４４では翼周りの循環が異なり、長翼４３に比べて短翼４４は循環の偏向力による流れの角度変化が小さくなり易い。
本実施形態では、短翼４４、長翼４３それぞれへの水の流入角度（βｗ＋Δβ）の相違を考慮して水車入口付近のキャンバーライン４５，４６を設定している。この結果、短翼４４、長翼４３での迎え角度の相違が小さくなり、キャビテーション性能が向上する。
【００５２】
（第５の実施の形態）
本発明の第５の実施形態を説明する。
本実施形態では、長翼４３の水車入口端角度βａ１を短翼４４の水車入口端角度βｂ１より大きくする。
羽根の水車入口端角度および翼長以外の条件は、長翼４３と短翼４４とで大きく変わることはない。即ち、本実施形態では、長翼４３と短翼４４の外径Ｄａ１、Ｄｂ１は等しいとしている。
これ以外の条件は第１の実施形態と本質的に異なる訳ではないので、全体の構成の図示、説明は省略する。
本実施形態において、長翼４３と短翼４４での迎え角の相違を低減することで、キャビテーション特性の向上を図っていることは、第１の実施形態と同様である。但し、本実施形態では、長翼４３と短翼４４の水車入口端角度に着目して長翼４３と短翼４４の迎え角の相違を低減している。
【００５３】
まず、ランナ４０の水車入口端における、羽根（長翼４３、短翼４４）に対する水の流れの平均的な相対流入角度φを表す式を導出する。この相対流入角度φは、迎え角に対応する量であり、長翼４３、短翼４４とで相対流入角度φが一致するように、長翼４３と短翼４４の水車入口端角度を設定する。
【００５４】
羽根の外径をＤ１［ｍ］、ランナ４０入口におけるのみ口高さをＢ［ｍ］、ランナ４０の回転速度をｎ［１／ｓｅｃ］、ランナ４０に流入する水の流量をＱ［ｍ^３／ｓｅｃ］、水の落差をＨ［ｍ］、ランナ４０に流入する水の平均径方向速度をＶｍ［ｍ／ｓｅｃ］、ランナ４０に流入する水の平均周方向速度をＶｔｈ［ｍ／ｓｅｃ］、ランナ４０に流入する水の流れの絶対角度をα［°］とする。
このとき、単位落差（√Ｈ）あたりの回転速度ｎ１は（ｎ１＝ｎ／√Ｈ）、単位落差（√Ｈ）あたりの流量Ｑ１は（Ｑ１＝Ｑ／√Ｈ）とそれぞれ表される。
【００５５】
図１６は、羽根の水車入口端での速度三角形を表す図である。この速度三角形により次の式（１）の関係がなりたつ。
Ｖｍ＝（Ｑ１・√Ｈ）／（π・Ｄ１・Ｂ）
ｕ＝π・Ｄ１・ｎ１・√Ｈ
Ｖｔｈ＝Ｖｍ／ｔａｎα …式（１）
【００５６】
相対流入角度φ（°）は、次の式（２）で表される。

【００５７】
長翼４３での単位落差あたりの回転速度ｎ１Ａ、流量Ｑ１Ａ、短翼４４での単位落差あたりの回転速度ｎ１Ｂ、流量Ｑ１Ｂを式（２）に代入する。
各運転点における長翼４３の相対流入角度φＡと、短翼４４の相対流入角度φＢの差（φＡ−φＢ）は、次の式（３）で表される。
【００５８】

【００５９】
回転速度ｎと流量Ｑを変化させてランナを動作させ、長翼４３と短翼４４それぞれにおける水車入口でのキャビテーションの初生点を測定した。
但し、ここでは羽根の長さ、形状が全て同一のランナ（スプリッタランナでない通常のフランシス形水車用のランナ）を用いた。これは、スプリッタランナの羽根（長翼、短翼）それぞれの近傍での流れの傾向と、通常のランナの羽根それぞれの近傍での流れの傾向とが大まかには一致する傾向があることによる。
即ち、ここでは、スプリッタランナで長翼、短翼の双方が存在することによる流れの複雑化要因を無視し、一種の近似化（あるいは、平均化）を行っている。
【００６０】
測定結果から相対的流入角度の差（φＡ−φＢ）を算出した結果、相対的流入角度の差（φＡ−φＢ）、即ち、長翼４３と短翼４４に対する水の流入角度の差は２°から６°の範囲であることが判った。
従って、この相対的流入角度差（φＡ−φＢ）に対応するように長翼４３と短翼４４の水車入口端角度β１をずらすことで、長翼４３と短翼４４に対する相対的な水の流入角度、即ち、互いの迎え角を近づけることができる。
【００６１】
具体的には、長翼４３の水車入口端角度βａ１と短翼４４の水車入口端角度βｂ１の差（βａ１−βｂ１）が算出された相対的流入角度差（φＡ−φＢ）に等しくなるように、長翼４３、短翼４４の水車入口端角度βａ１、βｂ１のいずれかまたは双方を修正する。例えば、短翼４４の水車入口端角度βｂ１を長翼４３の水車入口端角度βａ１に対してΔφＢ（＝φＡ−φＢ）だけずらす（２°≦ΔφＢ≦６°）。
このようにすることにより、流入する流れに対して長翼４３と短翼４４それぞれでの迎え角度の相違を小さくして、キャビテーション性能を向上することが可能となる。
【００６２】
（その他の実施形態）
本発明の実施形態は上記の実施形態に限られず拡張、変更可能であり、拡張、変更した実施形態も本発明の技術的範囲に含まれる。
上記実施形態では、クラウン、バンド間の円周方向に長翼４３と短翼４４とを交互に配置したスプリッタランナにおいて、長翼４３と短翼４４での流入角度に対する羽根角度（迎え角度）の差異を小さくすることで、キャビテーション性能等の向上を図っている。このように、羽根同士での迎え角度の相違を小さくすることが可能で有れば、本発明の実施形態に含まれる。
【００６３】
上記実施形態では、それぞれ単一のパラメータ（例えば、羽根の外径、厚み、バンド面との角度、反り、水車入口端角度）を長翼４３と短翼４４それぞれで調節することで、迎え角の均一化を図っていた。これに対して、複数のパラメータを任意に組み合わせることで迎え角の均一化を図ることも可能である。即ち、羽根の外径等のパラメータを２つ以上組み合わせても差し支えない。
【００６４】
【発明の効果】
以上説明したように、本発明によれば変落差運転時におけるキャビテーション性能の向上を図ったスプリッタランナおよび水力機械を提供できる。
【図面の簡単な説明】
【図１】本発明の第１実施形態に係る水力機械を側面から見た状態を表す一部断面図である。
【図２】本発明の第１実施形態に係る水力機械のランナを下方から見た状態を表す一部断面図である。
【図３】本発明の第１実施形態に係る水力機械のランナの子午面形状、および断面形状を表す図である。
【図４】本発明の実施形態の比較例たるランナの子午面形状、および断面形状を表す図である。
【図５】本発明の実施形態の比較例たるランナに流入する水の流れを表す図である。
【図６】本発明の第１実施形態に係る水力機械のランナのキャビテーション特性の１例を比較例と対比して表すグラフである。
【図７】本発明の第１実施形態に係る水力機械のランナの最高効率点付近での静圧分布の１例を比較例と対比して表すグラフである。
【図８】本発明の実施形態の比較例たるランナの高落差側、低落差側での静圧分布を表すグラフである。
【図９】本発明の第１の実施形態に係るランナの高落差側、低落差側での静圧分布の１例を表すグラフである。
【図１０】長翼と短翼の外径の比とキャビテーション無発生区間の関係の１例を表すグラフである。
【図１１】本発明の第２の実施形態に係るランナの断面形状を表す断面図である。
【図１２】長翼に対する短翼の羽根最大厚みの比と水力損失の関係の１例を表すグラフである。
【図１３】本発明の第３の実施形態に係るランナを側面から見た状態を表す側面図である。
【図１４】長翼の水車入口端曲線の角度とキャビテーションの関係の１例を表すグラフである。
【図１５】本発明の第４の実施の形態に係るランナの断面形状を表す断面図である。
【図１６】羽根の水車入口端での速度三角形を表す図である。
【符号の説明】
１０…水力機械、２０…ケーシング、３０…ガイドベーン、４０…ランナ、４１…クラウン、４２…バンド、４３…長翼、４４…短翼、５０…吸出管、６０…回転軸、７０…発電電動機[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a splitter runner and a hydraulic machine in which short blades are arranged between long blades.
[0002]
[Prior art]
In hydroelectric power generation (eg, pumped-storage power generation), a reaction water turbine that converts both water velocity energy and pressure energy into mechanical energy is used. Among the reaction water turbines, the Francis type water turbine has a runner in which a plurality of blades are arranged in the circumferential direction around the rotation axis, and the runner rotates by rotating water into the runner from the outer circumferential direction. Output is generated.
In Francis type turbines, runner blades are generally the same shape, but in order to improve turbine performance such as turbine efficiency, cavitation performance, low vibration, drop characteristics, etc., the blades are alternately long blades, A splitter runner with short blades may be used. In the splitter runner, the secondary flow between the blades is reduced due to the rectification effect due to the increase in the number of blades, and the hydraulic efficiency can be improved. Moreover, the cavitation performance can be improved by reducing the blade load per blade. In addition, the technique regarding a splitter runner is open | released (refer patent document 1, 2).
[0003]
[Patent Document 1]
JP 2000-54944 A
[Patent Document 2]
JP2000-205101A
[0004]
[Problems to be solved by the invention]
However, in the conventional splitter runner, the blade characteristics (for example, cavitation characteristics) are different between the long blades and the short blades.
For example, cavitation is likely to occur with long blades during high-head operation with a large water drop compared to the design point, and cavitation is likely to occur with short blades during low-head operation with a small water drop compared to the design point. In such a case, the range in which driving can be performed without cavitation is limited during the head change operation in which the head fluctuates.
In addition, the difference in characteristics between the long blade and the short blade may cause a reduction in efficiency in the low head region and the high head region.
In view of the above, it is an object of the present invention to provide a splitter runner and a hydraulic machine that improve the cavitation performance during a drop head operation.
[0005]
[Means for Solving the Problems]
(1) In order to achieve the above-mentioned object, the splitter runner according to the present invention includes a blade having a long blade length and a blade having a short blade length alternately in a circumferential direction around the rotation axis between the crown and the band. The at least a part of the end of the blade of the blade having a long blade length on the side of the inlet of the turbine between the bands, the inlet of the turbine of the blade having a short blade length, the inlet of the turbine between the bands. It is located on the inner diameter side of the circumference from the side end portion.
[0006]
The flow flowing into the runner applies a load to the short blade (short blade) located on the outer diameter side, and the flow with reduced pressure flows into the long blade (long blade). For this reason, compared with the case where the turbine blade entrance side ends of both the short blade and the long blade are located on the same circumference, the load that the short blade takes on increases and the load that the long blade takes on decreases. As a result, the blade loads of the long blades and the short blades are made uniform, the difference in deflection force acting on the turbine turbine inlet side ends of both blades is reduced, and the difference in angle of attack is reduced. As a result, cavitation characteristics can be improved.
Further, by shortening the long blades, friction loss is reduced as compared with a runner of the same size, and the turbine efficiency is improved.
[0007]
(2) A splitter runner according to the present invention is a splitter runner in which a blade having a long blade length and a blade having a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band. The maximum value Ta on the flow path with the blade length with the long blade length is smaller than the maximum value Tb on the flow path with the blade thickness with the short blade length.
[0008]
Since the maximum value Ta of the blades with short blades (short blades) is larger than the maximum value Tb of the blades with long blades (long blades), the circulation around the short blades is close to the circulation around the long blades. Therefore, the difference in blade load between the short blade and the long blade can be reduced. For this reason, a difference in deflection force acting at the turbine wheel inlet ends of both blades is reduced, a difference in angle of attack is reduced, and cavitation characteristics can be improved.
[0009]
(3) A splitter runner according to the present invention is a splitter runner in which a blade having a long blade length and a blade having a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band. The angle θa on the pressure surface side formed by the turbine inlet end curve along the turbine inlet side end of the blade having a long blade length and the surface of the band when viewed from the outer diameter direction of the circumference is the blade length. It is characterized by being smaller than the angle θb on the pressure surface side formed by the turbine blade inlet end curve of the short blade and the surface of the band.
[0010]
In general, during high-head operation, cavitation is more likely to occur near the suction-side turbine inlet end of a blade having a longer blade length (long blade) than at the suction-side turbine inlet end of a blade having a shorter blade length (short blade).
By making the angle θa of the turbine inlet end curve of the long blade smaller than the angle θb of the turbine inlet end curve of the short blade, the pressure drop near the suction side turbine inlet end of the long blade can be reduced compared to the short blade. . For this reason, the high head side cavitation limit of a long blade can be brought close to the high head side cavitation limit of a short blade having a larger limit value, and the cavitation limit during high head operation can be widened.
[0011]
(4) A splitter runner according to the present invention is a splitter runner in which a blade having a long blade length and a blade having a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band. The rate of change (∂βa / ∂r) of the angle βa with respect to the radius r of the circumference formed by the centerline of the blade with a long blade length and the circumference on the turbine inlet side is a blade with a short blade length. The angle βb formed between the center line and the circumference on the water turbine inlet side is greater than the rate of change (∂βb / ∂r) with respect to the radius r of the circumference.
[0012]
A difference in circulation angle around a long blade (long blade) and a circulation around a short blade (short blade) cause a difference in local inflow angle to each blade. By adjusting the center line of the blades near the water turbine inlet side, the difference in the angle of attack of both blades with respect to the inflow is reduced, and the cavitation characteristics can be improved.
[0013]
(5) A splitter runner according to the present invention is a splitter runner in which a blade having a long blade length and a blade having a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band. An angle βa on the suction surface side formed by the center line of the blade having a long blade length and the circumference on the inlet side of the turbine wheel, and a center line and the circumference of the blade having a short blade length on the turbine wheel inlet side. The difference (βa−βb) in the angle βb on the suction surface side formed by is 2 ° or more and 6 ° or less.
[0014]
By adjusting the turbine wheel inlet end angle corresponding to the difference in the local inflow angle between the long blade (long blade) and the short blade (short blade), the difference in the angle of attack of both blades The cavitation characteristics can be improved.
[0015]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.
(First embodiment)
FIG. 1 is a partial cross-sectional view showing a state in which a hydraulic machine 10 according to a first embodiment of the present invention is viewed from the side. FIG. 2 is a partial cross-sectional view showing a state in which the runner 40 of the hydraulic machine 10 is viewed from below (the suction pipe 50 side).
The hydraulic machine 10 includes a casing 20, a guide vane (guide vane) 30, a runner 40, a suction pipe 50, a rotating shaft 60, and a generator motor 70, and can be used for power generation (including pumped-storage power generation). That is, water flows in from the casing 20, passes through the guide vane 30 and the runner 40, and is discharged to the suction pipe 50, whereby the runner 40 is rotated to generate power.
The runner 40 is rotated by a water drop (the difference in water level between the upper pond and the lower pond in pumped storage power generation), and the hydraulic machine 10 functions as a so-called Francis type turbine.
[0016]
The casing 20 has a donut shape in which the cross-sectional area gradually decreases, and water flows from the upper pond during power generation.
The guide vane 30 adjusts the flow rate of water flowing into the runner 40 from the casing 20, and is arranged outside the runner 40 at a predetermined interval in the circumferential direction. Each guide vane 30 can rotate about its center, thereby adjusting the water flow rate. By the guide vanes 30, the water flow becomes a swirl flow having a circumferential velocity component and flows into the runner 40 from the outer peripheral direction.
[0017]
The runner 40 includes a crown 41, a band 42, long blades 43, and short blades 44, and is rotated by water flowing from the casing 20. The long blades 43 (blades having a blade length longer than the short blades 44) and the short blades 44 (blades having a blade length shorter than the long blades 43) are alternately arranged in the circumferential direction between the crown 41 and the band 42. That is, the runner 40 is a splitter runner. The detailed shape of the runner 40 (in particular, the long blades 43 and the short blades 44) will be described later.
[0018]
The suction pipe 50 allows the water used for the rotation of the runner 40 to flow out into a water discharge channel (a lower pond in the case of pumped-storage power generation).
The rotary shaft 60 transmits the rotary motion of the runner 40 to the generator motor 70.
The generator motor 70 generates power by rotating the rotary shaft 60.
[0019]
(Details of the shape of the runner 40)
3A and 3B are views showing the meridional shape and the cross-sectional shape of the runner 40, respectively. FIG. 3A corresponds to FIG. 1 and shows a shape of the runner 40 viewed from the meridian plane (a plane including the rotation axis 60). FIG. 3B shows a state in which the long blades 43 and the short blades 44 are cut along a plane F (a surface perpendicular to the rotation shaft 60) in FIG.
As shown in FIGS. 1 to 3, at least a part of the turbine inlet end curve La of the long blade 43 is located on the inner diameter side of the runner 40 with respect to the turbine inlet end curve Lb of the short blade 44.
[0020]
Here, the turbine inlet end curves La and Lb are curves along the turbine tip inlet side (casing 20 side) blade tip between the crown 41 and the band 42 for the long blade 43 and the short blade 44, respectively. Further, the turbine exit end curves Ea and Eb are curves along the blade tip on the turbine exit side (suction pipe 50 side) between the crown 41 and the band 42 for the long blade 43 and the short blade 44, respectively. Each of the circumferences Da1 and Db1 is an arc connecting a plurality of blade tips on the turbine inlet side of the long blade 43 and the short blade 44, and the rotation shaft 60 is the center thereof.
[0021]
(Basic idea in this embodiment)
The basic concept in this embodiment will be described.
FIGS. 4A and 4B are views showing a meridional shape and a cross-sectional shape of a runner 40x as a comparative example of the present embodiment, and correspond to FIGS. 3A and 3B.
In the runner 40x, although the long blades 43x and the short blades 44x have different blade lengths, their blade shapes are approximate, and the distance from the center of the rotating shaft 60 to the blade tip on the water turbine inlet side is equal. This is because the short wings 44x are not created separately from the long wings 43x, but are usually created by cutting the long wings 43x at a length of about 65 to 80% from the turbine inlet side. .
[0022]
FIG. 5 is a diagram illustrating the flow of water flowing into the runner 40x.
Water flows into the runner 40x from the upstream side (casing 20 side) at an inflow angle βw. In the vicinity of the blades (long blade 43x and short blade 44x), the inflow angle changes by Δβw to (βw + Δβw). This is due to the deflection force based on the circulation around each blade (long blade 43x and short blade 44x).
[0023]
Since the long blades 43 and the short blades 44 have different cross-sectional shapes, the blade loads are different. In general, since the blade load ΓA of the long blade 43 is larger than the blade load ΓB of the short blade 44, the change ΔβwA in the inflow angle near the long blade 43 is larger than the change ΔβwB in the inflow angle near the short blade 44 ( ΔβwA> ΔβwB). That is, the inflow angle (βw + ΔβwA) to the long blade 43 is larger than the inflow angle (βw + ΔβwB) to the short blade 44 (βw + ΔβwA> βw + ΔβwB).
[0024]
It has been found that the deterioration of the cavitation characteristics of the runner 40x occurs due to the difference in water inflow angle between the long blades 43 and the short blades 44. Hereinafter, the relationship between the difference in inflow angle between the long blades 43 and the short blades 44 and the cavitation characteristics of the runner 40x will be described.
[0025]
The difference in the inflow angle between the long blades 43 and the short blades 44 is the angle of attack (difference between the inflow angle and the blade angle, more precisely, the chord (the center line of the blade connecting the end of the turbine inlet and the end of the turbine outlet). ) Causes a difference in the direction of water flow (streamline).
When the angle of attack is large, the static pressure tends to decrease on the blade surface near the end of the blade on the turbine wheel entrance side. Accordingly, cavitation (a phenomenon in which the static pressure is locally lower than the vapor pressure and the water in the portion evaporates to generate water vapor bubbles) is likely to occur.
As a result, if the angle of attack differs between the long blades 43 and the short blades 44, the conditions for generating cavitation differ between the long blades 43 and the short blades 44, and cavitation tends to occur as a whole runner 40x. This is because even if cavitation occurs in either the long blades 43 or the short blades 44, cavitation occurs in the runner 40x.
[0026]
As described above, cavitation is likely to occur in the runner 40x due to the difference in water inflow angle (ultimate angle of attack) between the long blade 43 and the short blade 44. Conversely, by reducing the difference in the angle of attack between the long wings 43 and the short wings 44, the cavitation performance of the water turbine can be improved.
In the present embodiment, at least a part of the turbine inlet end curve La of the long blade 43 is positioned closer to the inner diameter side of the runner 40 than the turbine inlet end curve Lb of the short blade 44, thereby The aim is to eliminate the difference in angle of attack.
[0027]
(Characteristics of runner 40)
Hereinafter, the characteristics of the runner 40 according to the present embodiment will be described in comparison with the runner 40x of the comparative example.
FIGS. 6A and 6B are graphs showing cavitation characteristics in the runner 40x of the comparative example shown in FIG. 4 and the runner 40 of the present embodiment shown in FIGS. The horizontal axis of the graph is the water drop (the difference in water level between the upper pond and the lower pond), and the vertical axis represents the flow rate of water flowing into the runner.
A broken line represents an iso-efficiency curve connecting points with equal turbine efficiency. The solid line represents the initial point of cavitation at the turbine inlet of the long blade 43 (blade tip on the casing 20 side), and the alternate long and short dash line represents the initial point of cavitation at the turbine inlet of the short blade 44. Note that “x” represents the highest efficiency point at which the turbine efficiency is highest.
[0028]
7A and 7B are graphs showing the static pressure distribution in the vicinity of the maximum efficiency point of the runner 40x of the comparative example shown in FIG. 4 and the runner 40 of the present embodiment shown in FIGS. is there. 8A and 8B are graphs showing the static pressure distribution of the runner 40x of the comparative example on the high head side and the low head side, respectively. 9A and 9B are graphs showing the static pressure distribution of the runner 40 of the present embodiment on the high head side and the low head side, respectively.
7 to 9, the horizontal axis represents the length along the streamline on the blade surface, and the vertical axis represents the static pressure on the blade surface.
[0029]
As shown in FIG. 6A, in the comparative example, the initial cavitation point on the high head side is that the long blade 43x is on the lower head side than the short blade 44x. Conversely, the initial cavitation point on the low head side is higher on the short blade 44x than on the long blade 43x. For this reason, the cavitation non-occurrence section (the width of the head where cavitation does not occur at the flow rate at the highest efficiency point) S0 is narrow.
On the other hand, as shown in FIG. 6B, in this embodiment, the cavitation initial points of the long blades 43 and the short blades 44 coincide on both the high head side and the low head side. For this reason, the cavitation non-occurrence section (the width of the head where cavitation does not occur at the flow rate at the highest efficiency point) S1 becomes wide, and the cavitation characteristics at the turbine entrance are improved.
[0030]
As shown in FIGS. 8 and 9, the lowest point of the static pressure on the blade surface on the turbine wheel inlet side does not coincide between the long blade 43x and the short blade 44x in the comparative example, and the long blade 43 and the short blade in this embodiment. By matching at the wing 44. From FIG. 8, the static pressure at the tip of the turbine inlet of the long blade 43x on the high head side and the tip of the turbine blade of the short blade 44x on the low head side decreases below the saturated vapor pressure of water, and cavitation occurs. I understand that. 8 and 9 show the case of a high head and a low head, but there is no particular change in the tendency of the static pressure of the long blades 43 and the short blades 44 to coincide or disagree in the middle.
6 that the collision-free inflow points of the long blades 43 and the short blades 44 do not match in the comparative example, but also match in the present embodiment.
[0031]
In the runner 40 according to the present embodiment, a flow with a substantially constant pressure flowing in freely vortex from the upstream (casing 20 side) acts on the short blades 44 located on the outer diameter side of the runner 40, thereby shortening the flow. The wing 44 first has a load. Accordingly, a flow with reduced pressure acts on the long blades 43.
For this reason, in the runner 40 according to the present embodiment, compared to the comparative example in which the inlet ends of the short blades 44 and the long blades 43 are located on the same circumference, the load that the short blades 44 bear on increases, The load that the blades 43 are responsible for decreases. Accordingly, in the comparative example, the long blades 43x have a large load as shown in FIG. 7A and the short blades 44x do not have a large load, whereas in the present embodiment, as shown in FIG. 7B. Further, the blade loads of the long blades 43 and the short blades 44 are made uniform (the integrated value of the static pressure of the long blades 43 approaches the integrated value of the static pressure of the short blades 44).
[0032]
From the above, in this embodiment, the difference in the deflection force acting at the inlet side blade tips of the long blade 43 and the short blade 44 is reduced, and the difference in the flow inflow angle (βw + Δβw) is reduced. Accordingly, the difference in the pressure drop level at the inlet side blade tip during the drop head operation is also reduced.
As a result, as shown in FIG. 6B, the initial point of cavitation at the turbine entrance can be made substantially the same by the long blade 43 and the short blade 44, and the cavitation performance at the turbine entrance can be improved.
In addition, since the long blades 43 are shortened, the friction loss is reduced as compared with the runner 40x of the same size, so that the hydraulic efficiency is improved.
Furthermore, since the deviation in characteristics between the long blades 43 and the short blades 44 is improved and the collision-free inflow points coincide with each other, the maximum efficiency is improved.
[0033]
Next, the case where the outer diameters of the long blades 43 and the short blades 44 are changed will be quantitatively described.
FIG. 10 is a graph showing the relationship between the ratio of the outer diameters of the long blades 43 and the short blades 44 and the cavitation-free section S. The horizontal axis of the graph represents the ratio Da1 / Db1 between the outer diameter Da1 of the long blade 43 and the outer diameter Db1 of the short blade 44 in a cross section perpendicular to the rotation axis. The horizontal axis represents the ratio S1 / S0 of S to the cavitation-free section S1 with respect to the cavitation-free section S0 when the outer diameters of the long blade 43 and the short blade 44 are equal.
[0034]
When the ratio of the outer diameters of the long blades 43 and the short blades 44 (Da1 / Db1) decreases, the circulation ΓB around the short blades 44 increases and the circulation ΓA around the long blades 43 decreases. For this reason, the difference of the local inflow angle to each blade | wing becomes small, and S1 / S0 becomes large.
On the other hand, if Da1 / Db1 becomes too small, the short blades 44 have an excessive load, and the circulation ΓB around the short blades 44 exceeds the circulation ΓA around the long blades 43. For this reason, the positions of the short blades 44 and the long blades 43 are reversed, and S1 / S0 starts to decrease.
Therefore, as shown in FIG. 10, by setting the ratio Da1 / Db1 between the outer diameter of the long blade 43 and the outer diameter of the short blade 44 to a range of 0.85 or more and 0.98 or less (0.85 ≦ Da1 / Db1). ≦ 0.98), it is possible to increase the cavitation-free section S.
[0035]
(Second Embodiment)
A second embodiment of the present invention will be described.
FIG. 11 is a cross-sectional view showing a cross-sectional shape of a runner 40a according to the second embodiment of the present invention. This figure is different from FIG. 3B in that the long blades 43 and the short blades 44 are cut along a surface along the flow path of water from the turbine inlet (casing 20 side) to the turbine outlet (suction pipe 50 side), This represents the expanded state.
[0036]
In the present embodiment, the blade maximum thickness Ta of the long blades 43 is smaller than the blade maximum thickness Tb of the short blades 44 (Ta <Tb). The blade maximum thicknesses Ta and Tb are maximum values of the blade thickness when the blades (long blades 43 and short blades 44) are cut along a plane along the flow path.
Conditions other than the blade thickness and blade length do not change significantly between the long blade 43 and the short blade 44. That is, in the present embodiment, the outer diameters Da1 and Db1 of the long blades 43 and the short blades 44 are equal.
Since other conditions are not essentially different from those of the first embodiment, illustration and description of the entire configuration are omitted.
[0037]
In the present embodiment, the cavitation characteristics are improved by reducing the difference in angle of attack between the long blades 43 and the short blades 44, as in the first embodiment. However, in the present embodiment, the difference in angle of attack between the long blades 43 and the short blades 44 is reduced by paying attention to the thicknesses of the long blades 43 and the short blades 44.
By increasing the thickness of the short blade 44, the circulation around the short blade 44 is increased and the short blade 44 has a greater load. For this reason, the difference in blade load between the short blades 44 and the long blades 43 is reduced (uniformization of the blade load difference).
As a result, the difference in deflection force acting on the turbine blade inlet side blade tip of the long blade 43 and the short blade 44 is reduced, the difference in the attack angle is reduced, and the cavitation performance can be improved.
[0038]
Next, the relationship between the blade maximum thicknesses Ta and Tb of the long blades 43 and the short blades 44 and hydraulic loss will be described.
FIG. 12 is a graph showing an example of the relationship between the ratio Tb / Ta of the blade maximum thickness of the short blade 44 to the long blade 43 and hydraulic loss. The horizontal axis represents the ratio (Tb / Ta) of the blade maximum thickness Tb of the short blade 44 to the blade maximum thickness Ta of the long blade 43, and the vertical axis represents hydraulic loss. In this figure, the operating point where the inflow and the long blades 43 are matched is taken as an example.
[0039]
By increasing the circulation around the short blade 44, the angle change Δβw of the inflow is increased, and the collision loss of the short blade 44 is minimized in the vicinity of the collisionless inflow angle of the short blade 44. When the short blade 44 is made thicker and the circulation around the short blade 44 is increased, the collision loss of the short blade 44 starts to increase.
On the other hand, when the short blade 44 becomes thick, the cross-sectional area of the flow path decreases and the average flow velocity increases, so that the friction loss increases.
As a result, the hydraulic loss, which is the sum of collision loss and friction loss, decreases as the blade maximum thickness ratio Tb / Ta changes, and then increases.
[0040]
As shown in FIG. 12, the ratio Tb / Ta of the blade maximum thickness of the long blade 43 and the short blade 44 is 1.05 or more and 1.3 or less (1.05 ≦ Tb / Ta ≦ 1.3). Thus, the total hydraulic power loss could be minimized. In other words, the hydraulic loss of the short blade 44 can be reduced by making the short blade 44 thick to some extent.
In this case, by increasing the thickness of the short blade 44, it is possible to improve the strength of the short blade 44 together with the reduction of hydraulic loss.
[0041]
(Third embodiment)
A third embodiment of the present invention will be described.
FIG. 13: is a side view showing the state which looked at the runner 40b which concerns on the 3rd Embodiment of this invention from the side surface (outer diameter direction).
Long wings 43 and short wings 44 are alternately arranged in the circumferential direction between the crown 41 and the band 42. In the present embodiment, the angle θb formed between the turbine inlet end curve Lb of the short blade 44 and the surface of the band 42 is larger than the angle θa formed between the turbine inlet end curve La of the long blade 43 and the surface of the band 42.
[0042]
Precisely, the angle θa can be defined as the angle formed between the tangent line 6b and the tangent line 6c, and the angle θb can be defined as the angle formed between the tangent line 7b and the tangent line 7c.
The tangent line 6b is a tangent line of the turbine inlet end curve La at the intersection 6a between the turbine inlet end curve La of the long blades 43 and the band 42 as viewed from the outer diameter direction. The tangent 6c is a tangent at the intersection 6a opposite to the rotation direction of the circle passing through the intersection 6a and centering on the rotation axis 60.
The tangent line 7b is a tangent line of the turbine inlet end curve Lb at the intersection 7a between the turbine inlet end curve Lb of the short blade 44 and the band 42 as viewed from the outer diameter direction. The tangent line 7c is a tangent line at the intersection point 7a opposite to the rotation direction of the circle passing through the intersection point 7a and centering on the rotation axis 60.
[0043]
Conditions other than the angle of the turbine wheel end curve and the blade length do not change significantly between the long blade 43 and the short blade 44. That is, in the present embodiment, the outer diameters Da1 and Db1 of the long blades 43 and the short blades 44 are equal.
Since other conditions are not essentially different from those of the first embodiment, illustration and description of the entire configuration are omitted.
In the present embodiment, the difference in the angle of attack between the long blade 43 and the short blade 44 is reduced by paying attention to the angle of the turbine inlet end curve of the long blade 43 and the short blade 44.
[0044]
FIG. 14 is a graph showing an example of the relationship between the angle θa of the turbine inlet end curve of the long blades 43 and cavitation. The horizontal axis of the graph represents the angle θa of the turbine inlet end curve of the long blade 43. The vertical axis of the graph represents the minimum static pressure Psa-min (broken line graph) and hydraulic efficiency (solid line graph) on the suction surface of the long blades 43. Here, an operating point where cavitation is likely to occur on the suction surface near the water turbine inlet end is taken as an example.
[0045]
FIG. 14 shows that cavitation can be prevented without significantly reducing hydraulic efficiency by making the angle θb of the turbine inlet end curve of the short blade 44 larger than the angle θa of the turbine inlet end curve of the long blade 43 (static). Make water pressure higher than saturated vapor pressure).
This is because the angle θa at the long blades 43 is made smaller than the angle θb at the short blades 44, thereby mitigating the uneven flow at the suction surface near the water turbine inlet end band 42 (near the intersection 6a) of the long blades 43. By being done. The local pressure drop near the intersection 6a is suppressed by the relaxation of the flow bias.
[0046]
When the short blades 44 and the long blades 43 have the same blade shape on the water turbine inlet side, as shown in FIGS. 6 (A) and 8 (A), during the high head operation, the short blades 44 and the long blades 43 are more than the short blades 44. Cavitation occurs on the suction surface of the long blade 43 on the water turbine inlet side with a low head.
On the other hand, by making the angle θa of the long blades 43 smaller than the angle θb of the short blades 44, the pressure drop in the long blades 43 is reduced compared to the short blades 44, and the cavitation initial point of the long blades 43 is reduced. It can be moved to a higher head side. That is, the head where the cavitation occurs at the suction surface on the water turbine inlet side can be matched between the short blade 44 and the long blade 43, and the operating range on the high head side can be expanded.
[0047]
The angles θb and θa of the short blades 44 and the long blades 43 are set so that the high head cavitation characteristics at the turbine inlet satisfy the specifications. The angle θb of the short blade 44 is determined so that cavitation does not occur on the inlet suction surface of the short blade 44. Then, by increasing the angle difference (θb−θa) between the short blades 44 and the long blades 43 to some extent, the occurrence of cavitation on the inlet suction surface of the long blades 43 can be prevented.
On the other hand, as shown in FIG. 14, when the angle difference (θb−θa) is excessively increased, the hydraulic efficiency is lowered.
As shown in FIG. 14, by setting the angle difference (θb−θa) to 5 ° or more and 30 ° or less, a margin of hydrostatic pressure is secured with respect to saturated vapor pressure (pressure at which cavitation occurs), and hydraulic power is increased. It becomes possible not to impair the efficiency.
[0048]
(Fourth embodiment)
A fourth embodiment of the present invention will be described.
FIG. 15 is a cross-sectional view showing a cross-sectional shape of a runner 40c according to the fourth embodiment of the present invention. This figure represents the state which cut | disconnected the long blade 43 and the short blade 44 in the surface along the flow path of the water from a turbine inlet (casing 20 side) to a turbine exit (suction pipe 50 side), and expand | deployed this.
[0049]
In the present embodiment, the warp amount Ca of the long blades 43 is larger than the warp amount Cb of the short blades 44 on the water turbine inlet side.
Here, the warp amount Ca of the long blades 43 is determined by the camber line 45 of the long blades 43 (the center line of the long blades 43 connecting the water turbine inlet side end and the water turbine outlet side end along the flow path) and the rotation axis. It can be defined as a rate of change (∂βa / ∂r) with respect to the radius r of the angle βa formed by a circle of radius r centered at 60.
Further, the amount of warpage Cb of the short blade 44 is equal to the camber line 46 of the short blade 44 (the center line of the short blade 44 connecting the water turbine inlet side end and the water turbine outlet side end along the flow path) and the radius r. It can be defined as the rate of change (∂βb / ∂r) with respect to the radius r of the angle βb with the circle.
[0050]
Conditions other than the amount of blade warpage and the blade length do not change significantly between the long blade 43 and the short blade 44. That is, in the present embodiment, the outer diameters Da1 and Db1 of the long blades 43 and the short blades 44 are equal.
Since other conditions are not essentially different from those of the first embodiment, illustration and description of the entire configuration are omitted.
[0051]
In the present embodiment, the cavitation characteristics are improved by reducing the difference in angle of attack between the long blades 43 and the short blades 44, as in the first embodiment. However, in the present embodiment, the difference in angle of attack between the long blades 43 and the short blades 44 is reduced by paying attention to the amount of warpage between the long blades 43 and the short blades 44.
As shown in FIG. 5, the circulation around the blades is different between the long blades 43 and the short blades 44, and the short blades 44 tend to have a smaller change in the flow angle due to the deflection force of the circulation than the long blades 43.
In the present embodiment, the camber lines 45 and 46 in the vicinity of the turbine inlet are set in consideration of the difference in the inflow angle (βw + Δβ) of water into the short blade 44 and the long blade 43, respectively. As a result, the difference in angle of attack between the short blade 44 and the long blade 43 is reduced, and the cavitation performance is improved.
[0052]
(Fifth embodiment)
A fifth embodiment of the present invention will be described.
In the present embodiment, the turbine inlet end angle βa1 of the long blades 43 is made larger than the turbine inlet end angle βb1 of the short blades 44.
Conditions other than the turbine wheel inlet end angle and the blade length of the blade do not change significantly between the long blade 43 and the short blade 44. That is, in the present embodiment, the outer diameters Da1 and Db1 of the long blades 43 and the short blades 44 are equal.
Since other conditions are not essentially different from those of the first embodiment, illustration and description of the entire configuration are omitted.
In the present embodiment, the cavitation characteristics are improved by reducing the difference in angle of attack between the long blades 43 and the short blades 44, as in the first embodiment. However, in the present embodiment, the difference in the angle of attack between the long blades 43 and the short blades 44 is reduced by paying attention to the water turbine inlet end angle of the long blades 43 and the short blades 44.
[0053]
First, an equation representing the average relative inflow angle φ of the water flow with respect to the blades (long blade 43, short blade 44) at the turbine wheel inlet end of the runner 40 is derived. The relative inflow angle φ is an amount corresponding to the angle of attack, and the turbine inlet end angle of the long blade 43 and the short blade 44 is set so that the relative inflow angle φ matches between the long blade 43 and the short blade 44. .
[0054]
The outer diameter of the blade is D1 [m], the opening height at the inlet of the runner 40 is B [m], the rotational speed of the runner 40 is n [1 / sec], and the flow rate of water flowing into the runner 40 is Q [m. ³ / Sec], the head of water is H [m], the average radial velocity of water flowing into the runner 40 is Vm [m / sec], and the average circumferential velocity of water flowing into the runner 40 is Vth [m / sec]. The absolute angle of the flow of water flowing into the runner 40 is α [°].
At this time, the rotational speed n1 per unit head (√H) is expressed as (n1 = n / √H), and the flow rate Q1 per unit head (√H) is expressed as (Q1 = Q / √H).
[0055]
FIG. 16 is a diagram illustrating a speed triangle at the turbine wheel inlet end of the blade. The relationship of the following formula (1) is established by this speed triangle.
Vm = (Q1 · √H) / (π · D1 · B)
u = π · D1 · n1 · √H
Vth = Vm / tan α (1)
[0056]
The relative inflow angle φ (°) is expressed by the following equation (2).

[0057]
The rotational speed n1A per unit head at the long blade 43, the flow rate Q1A, the rotational speed n1B per unit head at the short blade 44, and the flow rate Q1B are substituted into the equation (2).
The difference (φA−φB) between the relative inflow angle φA of the long blades 43 and the relative inflow angle φB of the short blades 44 at each operating point is expressed by the following equation (3).
[0058]

[0059]
The runner was operated by changing the rotational speed n and the flow rate Q, and the initial point of cavitation at the turbine entrance at each of the long blade 43 and the short blade 44 was measured.
However, here, a runner having the same blade length and shape (a normal runner for Francis type turbines that is not a splitter runner) was used. This is because the tendency of the flow in the vicinity of each of the blades of the splitter runner (long blade and short blade) and the tendency of the flow in the vicinity of each of the blades of the normal runner tend to roughly match.
That is, in this case, a kind of approximation (or averaging) is performed by ignoring the complicating factor of the flow due to the presence of both long blades and short blades in the splitter runner.
[0060]
As a result of calculating the relative inflow angle difference (φA−φB) from the measurement result, the relative inflow angle difference (φA−φB), that is, the difference in water inflow angle between the long blade 43 and the short blade 44 is 2 °. It was found to be in the range of 6 °.
Accordingly, the water inflow relative to the long blade 43 and the short blade 44 is shifted by shifting the turbine inlet end angle β1 of the long blade 43 and the short blade 44 so as to correspond to this relative inflow angle difference (φA−φB). The angles, that is, the angles of attack of each other can be made closer.
[0061]
Specifically, the difference (βa1−βb1) between the turbine inlet end angle βa1 of the long blade 43 and the turbine inlet end angle βb1 of the short blade 44 is equal to the calculated relative inflow angle difference (φA−φB). One or both of the turbine inlet end angles βa1 and βb1 of the long blades 43 and the short blades 44 are corrected. For example, the turbine inlet end angle βb1 of the short blade 44 is shifted by ΔφB (= φA−φB) with respect to the turbine inlet end angle βa1 of the long blade 43 (2 ° ≦ ΔφB ≦ 6 °).
By doing so, it is possible to improve the cavitation performance by reducing the difference in angle of attack between the long blades 43 and the short blades 44 with respect to the inflowing flow.
[0062]
(Other embodiments)
Embodiments of the present invention are not limited to the above-described embodiments, and can be expanded and modified. The expanded and modified embodiments are also included in the technical scope of the present invention.
In the above embodiment, in the splitter runner in which the long blades 43 and the short blades 44 are alternately arranged in the circumferential direction between the crown and the band, the blade angle (attack angle) with respect to the inflow angle between the long blades 43 and the short blades 44 is set. By reducing the difference, the cavitation performance is improved. Thus, if it is possible to reduce the difference in the angle of attack between the blades, it is included in the embodiment of the present invention.
[0063]
In the above-described embodiment, the angle of attack is adjusted by adjusting the single parameters (for example, the outer diameter of the blade, the thickness, the angle with the band surface, the warp, and the turbine inlet end angle) with the long blade 43 and the short blade 44, respectively. Was made uniform. On the other hand, it is possible to make the angle of attack uniform by arbitrarily combining a plurality of parameters. That is, two or more parameters such as the outer diameter of the blade may be combined.
[0064]
【The invention's effect】
As described above, according to the present invention, it is possible to provide a splitter runner and a hydraulic machine that improve the cavitation performance during a drop head operation.
[Brief description of the drawings]
FIG. 1 is a partial cross-sectional view showing a state of a hydraulic machine according to a first embodiment of the present invention as viewed from the side.
FIG. 2 is a partial cross-sectional view showing a state in which the runner of the hydraulic machine according to the first embodiment of the present invention is viewed from below.
FIG. 3 is a diagram showing a meridional shape and a cross-sectional shape of a runner of the hydraulic machine according to the first embodiment of the present invention.
FIG. 4 is a diagram illustrating a meridian shape and a cross-sectional shape of a runner as a comparative example of the embodiment of the present invention.
FIG. 5 is a diagram showing a flow of water flowing into a runner as a comparative example of the embodiment of the present invention.
FIG. 6 is a graph showing an example of the cavitation characteristics of the runner of the hydraulic machine according to the first embodiment of the present invention in comparison with a comparative example.
FIG. 7 is a graph showing an example of a static pressure distribution in the vicinity of the highest efficiency point of the runner of the hydraulic machine according to the first embodiment of the present invention in comparison with a comparative example.
FIG. 8 is a graph showing a static pressure distribution on a high head side and a low head side of a runner as a comparative example of the embodiment of the present invention.
FIG. 9 is a graph showing an example of a static pressure distribution on a high head side and a low head side of the runner according to the first embodiment of the present invention.
FIG. 10 is a graph showing an example of the relationship between the ratio of the outer diameter of the long blade and the short blade and the cavitation-free section.
FIG. 11 is a cross-sectional view illustrating a cross-sectional shape of a runner according to a second embodiment of the present invention.
FIG. 12 is a graph showing an example of the relationship between the ratio of the maximum blade thickness of the short blade to the long blade and hydraulic loss.
FIG. 13 is a side view showing a state in which a runner according to a third embodiment of the present invention is viewed from the side.
FIG. 14 is a graph showing an example of the relationship between the angle of the turbine blade inlet end curve of long blades and cavitation.
FIG. 15 is a cross-sectional view showing a cross-sectional shape of a runner according to a fourth embodiment of the present invention.
FIG. 16 is a diagram showing a speed triangle at a turbine wheel inlet end of a blade.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 10 ... Hydraulic machine, 20 ... Casing, 30 ... Guide vane, 40 ... Runner, 41 ... Crown, 42 ... Band, 43 ... Long wing, 44 ... Short wing, 50 ... Suction pipe, 60 ... Rotating shaft, 70 ... Generator motor

Claims

A splitter runner in which a blade with a long blade length and a blade with a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band,
The crown of the blade with a long blade length, at least a part of the end portion on the water turbine inlet side between the bands, and the inner diameter of the circumference with respect to the crown of the blade with a short blade length, the end portion on the water wheel inlet side between the bands. Splitter runner characterized by being located on the side.

The ratio Da1 / Db1 between the outer diameter Da1 of the turbine blade inlet side end of the long blade and the outer diameter Db1 of the turbine blade inlet end of the short blade is 0.85 or more and 0.98 or less. The splitter runner according to claim 1, wherein:

A splitter runner in which a blade with a long blade length and a blade with a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band,
A splitter runner characterized in that a maximum value Ta on a flow path with a long blade length is smaller than a maximum value Tb on a flow path with a short blade length.

The ratio Tb / Ta between the maximum thickness Tb of the blade having a short blade length and the maximum value Ta of the blade having a long blade length is 1.05 or more and 1.3 or less. Item 4. A splitter runner according to item 3.

A splitter runner in which a blade with a long blade length and a blade with a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band,
The angle θa on the pressure surface side formed by the turbine inlet end curve along the turbine inlet side end of the blade having a long blade length and the surface of the band as viewed from the outer diameter direction of the circumference is a blade having a short blade length. The splitter runner is characterized in that it is smaller than the angle θb on the pressure surface side formed by the water turbine inlet end curve and the surface of the band.

6. The splitter runner according to claim 5, wherein a difference (θb−θa) between the angle θb of the blade having the short blade length and the angle θa of the blade having the long blade length is 5 ° or more and 30 ° or less.

A splitter runner in which a blade with a long blade length and a blade with a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band,
The rate of change (∂βa / ∂r) of the angle βa between the centerline of the blade with a long blade length and the circumference on the turbine inlet side with respect to the radius r of the circumference is the centerline of the blade with a short blade length. And the circumference is greater than the rate of change (∂βb / ∂r) of the angle βb with respect to the radius r of the circumference at the turbine inlet side.

A splitter runner in which a blade with a long blade length and a blade with a short blade length are alternately arranged in a circumferential direction around a rotation axis between a crown and a band,
An angle βa on the suction surface side formed by the center line of the long blade and the circumference on the turbine inlet side, and a negative angle formed by the center line of the short blade and the circumference on the turbine entrance side. A splitter runner characterized in that a pressure-side angle βb difference (βa−βb) is 2 ° or more and 6 ° or less.

A hydraulic machine comprising the splitter runner according to claim 1.