JP4464629B2

JP4464629B2 - Centrifugal pump for liquid

Info

Publication number: JP4464629B2
Application number: JP2003172638A
Authority: JP
Inventors: 銀春曹; 秀樹一井
Original assignee: Sanso Electric Co Ltd
Current assignee: Sanso Electric Co Ltd
Priority date: 2003-06-17
Filing date: 2003-06-17
Publication date: 2010-05-19
Anticipated expiration: 2023-06-17
Also published as: JP2005009361A

Description

【０００１】
【発明の属する技術分野】
本発明は、遠心流体機械に有する遠心羽根車の構造のうち、特に液体用の遠心ポンプの遠心羽根車の構造に関する。
【０００２】
【従来技術】
従来、図１に示す液体用の遠心ポンプの遠心羽根車（以下、「羽根車」という。）１では、図２に示すように、羽根車の羽根圧力面７から羽根負圧面８までの羽根間１２、あるいは羽根厚さを無視した場合での、羽根圧力面から羽根負圧面までの羽根間のＺ×θ［θ＝３６０°／Ｚ、Ｚ：羽根枚数］の範囲で、かつ、羽根車入口１４から羽根車出口１５に向っての各羽根半径において、軸方向流路幅１６が一定である。ところで、角速度ωで回転する羽根車１内を総体速度Ｗで流れる密度ρの流体には、単位面積当たり２ρωＷのコリオリ力が他の力より大きく作用する。一方、小形低比速度の液体用の遠心ポンプに対しては、軸方向流路幅１６が狭くなり、２ρωＷのコリオリ力もさらに大きくなると考えられる。このため、図３の（ｂ）に示すように、羽根車主板４の面または羽根車側板６の面の近くの境界層では、流体は羽根５間の羽根圧力面７から羽根負圧面８に向って流れて、羽根負圧面８付近、特に羽根車側板６の面と羽根負圧面８の角の付近に低エネルギー流体２０が集積し、この低エネルギー流体２０である二次流れが流路通路の大部分を占め、羽根負圧面８付近における主流１９である非低エネルギー流体の有効流路幅が著しく狭くなる。このため、羽根５間で圧力が一様でなくなり、図１において、羽根車出口１５の流れと渦巻室２の舌部３などとのポテンシャル干渉と粘性後流干渉によって、吐出口１０における圧力は、（回転数ｎ）×（羽根枚数ｚ）の基本成分を持つ周期変動ΔＰ（圧力脈動）を繰り返す。
【０００３】
液体用の遠心ポンプの低振動化と低騒音化のため、従来、図１に示す、渦巻室２の舌部３と羽根車外径１１との隙間１８を拡大したり、渦巻室２の舌部３の形状を変えるなどの対策が行われてきた。このようにすることで、液体用の遠心ポンプの振動および騒音または圧力脈動はある程度低減することができる。しかし、羽根車１と渦巻室２との間における流体損失が多く生じる場合があり、液体用の遠心ポンプの性能が下がってしまう傾向もある。
【０００４】
一方、液体用の遠心ポンプの羽根車１を遠心ポンプ（以下、「ポンプ」という。）の羽根車１とした場合では、ポンプのキャビテーション性能については、羽根車１の羽根枚数を減らしたり、羽根車入口１４における流速を遅くしたり、羽根５の形状を最適化すること等によって、ポンプのキャビテーション性能を向上できる。しかし、ポンプのキャビテーション性能とポンプの効率との間にトレードオフの関係があり、これらがポンプの高効率維持に悪影響をもたらす問題がある。
【０００５】
【特許文献１】
特願２００２−１１９７６４（平成１４年４月２２日出願）
【０００６】
【発明が解決しようとする課題】
本発明では、液体用の遠心ポンプのイニシャルコストアップ無しに、液体用の遠心ポンプの高効率またはポンプとした場合の高キャビテーション性能を維持して、羽根車出口流れと渦巻室の舌部とのポテンシャル干渉と粘性後流干渉の低減を図ることによって、圧力脈動の低減化および騒音の低減化を実現できる液体用の遠心ポンプを提供することである。
【０００７】
【課題を解決するための手段】
上記の課題を解決するための本発明の手段は、モータにより駆動されることにより、流体を吸込口９に吸込み、渦巻室２内の羽根車１から遠心力により羽根車１の周辺の渦巻室２の周辺部へ吐出し、吐出された流体を渦巻室２の周辺部から集めて吐出口１０へ一括して吐出する吐出管１３からなる液体用の遠心ポンプであり、この液体用の遠心ポンプでは羽根車１は最も重要な要素であって、複数枚の羽根５とそれらが結合されている羽根車主板４と羽根車側板６からなり、羽根車１の隣り合う２枚の羽根５並びに羽根車主板４と羽根車側板６で形成される羽根間流路１２は、羽根車入口１４から羽根車出口１５に向かう間に、羽根圧力面７側の軸方向流路幅１６を羽根負圧面８側の軸方向流路幅１７と相違するものとした遠心機械に関するものである。この場合、羽根圧力面７から羽根負圧面８までの羽根５間で、かつ、羽根車入口１４から羽根車出口１５に向かう間に、羽根圧力面７側の軸方向流路幅１６と羽根負圧面８側の軸方向流路幅１７が相違して行くものも上記の遠心粒体機械に含まれる。
【０００８】
このようにすることによって、液体用の遠心ポンプの運転時に、羽根車１の、羽根間流路１２における流れまたは遠心力とコリオリ力などにより生じた低エネルギー流体２０の二次流れと羽根車出口１５の流れ分布をコントロールすることを可能とする。
【０００９】
すなわち、第１の手段では、遠心羽根車を有する液体用の遠心ポンプにおいて、羽根車１は、羽根車出口１５側における羽根圧力面７から羽根負圧面８までの羽根５間で、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が拡大し、さらに羽根入口１４側における羽根圧力面７から羽根負圧面８までの羽根５間で、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が拡大し、かつ、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が拡大した状態を保ちつつ羽根車入口１４から羽根車出口１５まで形成されていることを特徴とする液体用の遠心ポンプである。
第２の手段では、遠心羽根車を有する液体用の遠心ポンプにおいて、羽根車１は、羽根車出口１５側における羽根圧力面７から羽根負圧面８までの羽根５間で、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が拡大し、さらに羽根入口１４側における羽根圧力面７から羽根負圧面８までの羽根５間で、羽根圧力面７側の軸方向流路幅１６が羽根負圧面８側の軸方向流路幅１７と同じで、かつ、羽根圧力面７側の軸方向流路幅１６と羽根負圧面８側の軸方向流路幅１７と同じである状態の羽根車入口１４から羽根車出口１５になるに連れて羽根圧力面７側と羽根負圧面８側の間の軸方向流路幅が徐々に縮小すると共に、羽根圧力面７側の軸方向流路幅１６が羽根負圧面８側の軸方向流路幅１７より縮小して形成されていることを特徴とする液体用の遠心ポンプである。
【００１０】
このようにすることで羽根負圧面８付近における主流の有効流路幅が増やされ、羽根車出口１５における流れの周方向分布がある程度均一化されることが考えられ、羽根車出口１５の流れと渦巻室２の舌部３との干渉を低減することで、液体用の遠心ポンプの圧力脈動や騒音を低減できる。
【００１１】
一方、羽根車入口１４においても、羽根負圧面８側の羽根車主板４と羽根車側板６の間である軸方向流路幅１７は羽根圧力面７側の羽根車主板４と羽根車側板６の間である軸方向流路幅１６と同じであるか、軸方向流路幅１６に比べて徐々に拡大されているものである。このようにすることで、液体用の遠心ポンプの羽根車１、例えば遠心ポンプの羽根車１としたときのポンプの高キャビテーション性能を維持することができる。
【００１２】
第３の手段では、羽根車１は、外径が２００ｍｍ以下で、かつ、比速度Ｎｓ＝ｎＱ^１／２／Ｈ^３／４が２００以下の小型低比速度であることを特徴とする上記の第１または第２の手段の液体用の遠心ポンプである。
【００１３】
第４の手段では、羽根車１は、その最大外径における対面する２枚の羽根５、５間において、羽根圧力面７側の軸方向流路幅１６に対する羽根負圧面８側の軸方向流路幅１７が１より大で、かつ、４以下であることを特徴とする第１〜３の手段のいずれか１の手段の液体用の遠心ポンプである。
【００１４】
ところで、一般的に液体用の遠心ポンプの性能の面から考慮して、羽根車入口１４における軸方向流路幅は羽根車出口１５における軸方向流路幅と同じとするか又はそれより大きくすることを基本とする。その上で、羽根車入口１４側における羽根圧力面７側の軸方向流路幅１６に対する羽根負圧面８側の軸方向流路幅１７の比率を異なるものとする場合もある。一方、圧力脈動の低減のみを重視する場合では、羽根車入口１４側において、従来の様に羽根圧力面７側の軸方向流路幅１６に対する羽根負圧面８側の軸方向流路幅１７を同一とすることもある。
【００１５】
ここで、本発明の作用について説明する。先ず、従来の液体用の遠心ポンプの低振動化や低騒音化に当たっては、渦巻室２の舌部３と羽根車外径１１との隙間１８を広げたり、渦巻室２の舌部３の形状を変えることなどによって、液体用の遠心ポンプの振動および騒音をある程度まで低減できた。しかし、これらは、羽根車１と渦巻室２との間における流体損失が多く生じる場合があり、液体用の遠心ポンプの性能が下がってしまう傾向があり、また液体用の遠心ポンプのサイズアップを図らなければならない傾向もある。一方、液体用の遠心ポンプとした場合のポンプのキャビテーション性能については、羽根車１の羽根５の枚数を減らしたり、羽根車入口１４における流速を遅くさせたり、羽根車入口１４の形状と羽根５の枚数を減らしたり、羽根車入口１４における流速を遅くさせたり、羽根５の形状を最適化することなどによって、キャビテーション性能が向上できる。しかし、ポンプのキャビテーション性能とポンプの効率との間にはトレードオフの関係があり、ポンプの高効率維持に悪影響をもたらす。これに対して、本発明の流体機械では、請求項１の手段における羽根車１を採用することによって液体用の遠心ポンプの高効率、かつ、遠心ポンプとした場合のポンプの高キャビテーション性能を維持することができ、圧力変動が大幅に低減される。
【００１６】
【発明の実施の形態】
本発明の実施の形態を図面を参照して説明する。図１は羽根車と渦巻室の断面図である。図２は従来の液体用の遠心ポンプの羽根車の構造をその内径側（ａ）と外径側（ｂ）で切断した端面を模式的に直線化して示して説明する断面図と、羽根車の半径分を示す模式図である。図３は、従来の羽根車出口における流体の流れ構造および圧力脈動の説明図て、（ａ）は圧力脈動の模式図で、（ｂ）は羽根車内の流れ構造の模式図である。図４は本発明の液体用の遠心ポンプの一実施の形態の羽根車の構造をその内径側（ａ）と外径側（ｂ）で切断した端面を模式的に直線化して示して説明する断面図と、羽根車の半径分を示す模式図である。図５は本発明の液体用の遠心ポンプの他の一実施の形態の羽根車の構造をその内径側（ａ）と外径側（ｂ）で切断した端面を模式的に直線化して示して説明する断面図と、羽根車の半径分を示す模式図である。図６は本発明の液体用の遠心ポンプのさらに他の一実施の形態の羽根車の構造をその内径側（ａ）と外径側（ｂ）で切断した端面を模式的に直線化して示して説明する断面図と、羽根車の半径分を示す模式図である。図７は本発明の液体用の遠心ポンプのさらに他の一実施の形態の羽根車の構造をその内径側（ａ）と外径側（ｂ）で切断した端面を模式的に直線化して示して説明する断面図と、羽根車の半径分を示す模式図である。図８は従来の羽根車の遠心ポンプと本発明の羽根車の遠心ポンプの性能比較図である。図９は従来の羽根車の遠心ポンプと本発明の羽根車のＮＰＳＨ３（３％揚程低下）の測定比較図である。図１０は従来の羽根車の圧力脈動波形を示す図である。図１１は本発明の羽根車の圧力脈動波形を示す図である。
【００１７】
本発明の一実施の形態は液体用の遠心ポンプである。図１に示すように、この遠心ポンプは、渦巻室２内に、モータの駆動により汲み上げる液体を、遠心ポンプの吸込口９に吸込み、遠心ポンプの吐出口１０に吐出すための、羽根車１を有し、さらに羽根車１より吐出され液体を集める渦巻室２を有する。この羽根車１は本発明の液体用の遠心ポンプにおいて最も重要な要素であって、複数枚の羽根５と羽根５を垂直に結合して挟持している羽根車主板４と羽根車側板６から形成されている。羽根車１の隣り合う２枚の羽根５並びに羽根車主板４と羽根車側板６で形成される羽根間流路１２は、図４に示すように、羽根負圧面８側の軸方向流路幅１７が羽根圧力面７側の軸方向流路幅１６に比べて徐々に拡大されている状態を保ちつつ、羽根車入口１４から羽根車出口１５まで形成されている。
【００１８】
このようにすることで、羽根車１の羽根負圧面８付近の流れ有効流路を拡大して、羽根車出口１５側における流れを含めた羽根間流路１２の流れがある程度均一化され、羽根車出口１５の流れと渦巻室２の舌部３との干渉を低減することで、液体用の遠心ポンプの圧力脈動や騒音を低減できる。一方、羽根車１の羽根車入口１４においては、羽根負圧面８側の軸方向流路幅１７が羽根圧力面７側の軸方向流路幅１６に比べ徐々に拡大され、羽根車入口１４側付近の羽根負圧面８における相対流速も減少されることによって、遠心ポンプの羽根車１としたポンプの高キャビテーション性能を維持できる。
【００１９】
さらに、図５に示す実施の形態では、その（ａ）に示すように羽根車入口１４側である羽根車１の内径側の部分では、軸方向流路幅は羽根圧力面７側が羽根車負圧面８側も共に同じ軸方向流路幅を有している。しかし、図５の（ｂ）に示すように羽根車出口１５側である羽根車１の外径側の部分では、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が徐々に拡大されている。
【００２０】
さらに、図６に示す実施の形態では、図５における実施の形態と同様に、その（ａ）に示すように、羽根車入口１４側である羽根車１の内径側の部分では、軸方向流路幅は羽根圧力面７側が羽根負圧面８側も共に同じ軸方向流路幅である。しかし、図６の（ｂ）に示すように羽根車出口１５側である羽根車１の外径側の部分では、羽根車主板４の板厚は一定のものであるが、羽根車側板６の板厚はその内面側のみにテーパーが形成されることで、羽根圧力面７側の軸方向流路幅１６に比して羽根負圧面８側の軸方向流路幅１７が徐々に拡大されている。
【００２１】
さらに、図７に示す実施の形態では、その（ａ）に示すように羽根車入口１４側である羽根車１の内径側の部分では、軸方向流路幅は羽根圧力面７側が羽根負圧面８側も共に同じ軸方向流路幅である。しかし、図７の（ｂ）に示すように羽根車出口１５側である羽根車１の外径側の部分では、羽根圧力面７側の軸方向流路幅１６が羽根負圧面８側への一定の長さの間は軸方向流路幅１６が一定の狭い幅をしているが、その途中から羽根負圧面８側になるに連れて羽根負圧面８側の軸方向流路幅１７が徐々に拡大されて行くものである。
【００２２】
以上のようにしたことで、本発明の液体用の遠心ポンプの実施の形態である遠心ポンプでは、低圧力脈動や低騒音とすることができる。
【００２３】
さらに、他の実施の形態では、羽根圧力面７側の軸方向流路幅１６に対する羽根負圧面８側の軸方向流路幅１７は、１より大で、かつ、４以下とする小形低比速度の遠心ポンプである。
【００２４】
【実施例】
先ず、図４に示す本発明の実施例の遠心ポンプの仕様の諸寸法を例示すると、羽根車１の外径は６２ｍｍ、羽根車入口１４の羽根負圧面８の軸方向流路幅は５．８ｍｍ、羽根車入口１４の羽根圧力面７の軸方向流路幅は３．８ｍｍ、羽根車出口１５の羽根負圧面８の軸方向流路幅は４．５ｍｍ、羽根車出口１５の羽根圧力面７の軸方向流路幅は２．５ｍｍである。
【００２５】
さらに、図５に示す本発明の実施例の遠心ポンプの仕様の諸寸法を例示すると、羽根車１の外径は６２ｍｍ、羽根車入口１４の軸方向流路幅は羽根負圧面８も羽根圧力面７も同じ幅で４．８ｍｍ、羽根車出口１５の羽根負圧面８の軸方向流路幅は４．５ｍｍ、羽根車出口１５の羽根圧力面７の軸方向流路幅は２．５ｍｍである。
【００２６】
なお、比較のために従来の軸方向流路幅が羽根負圧面側も羽根圧力面側も同一である遠心ポンプの諸寸法を示すと、羽根車１の外径は６２ｍｍ、羽根車入口１４の軸方向流路幅は４．８ｍｍ、羽根車出口１５の軸方向流路幅は３．５ｍｍである。上記において、渦巻室２を含めた遠心ポンプの他の諸寸法は全て同一であり、設計流量Ｑ_dは３０ｌ／ｍｉｎとしている。
【００２７】
図８に従来の羽根車を装備した遠心ポンプと本発明における上記の図４に示す実施例の羽根車１を装備した遠心ポンプの性能を比較して示す。ただし、本発明の図５に示す実施例の羽根車１を装備した遠心ポンプの性能が図４に示す実施例の羽根車１を装備した遠心ポンプの性能一致しているため、本発明における図４に示す羽根車１を装備した遠心ポンプの性能を従来の羽根車を装備した遠心ポンプと比較している。全流量域において、本発明における羽根車を装備した遠心ポンプ効率は、従来のポンプの効率と変わらないものの、全揚程Ｈが少々高く現れ、遠心羽根車の内部流れと出口流れがコントロールされていることが推測される。
【００２８】
図９に、従来の羽根車を装備した小形低比速度遠心ポンプと本発明における上記の図４の実施例の羽根車１を装備した小形低比速度遠心ポンプの揚程３％低下を対比して示す。従来のポンプに比べ、揚水量Ｑが増加するに連れて本発明の遠心羽根車を有する遠心ポンプのキャビテーション性能が僅かに良くなっている傾向が見られる。
【００２９】
図１０に、従来の羽根車を装備した遠心ポンプの吐出口における圧力変動を示し、図１１に本発明の図４に示す実施例の羽根車を装備した遠心ポンプの吐出口における圧力変動を、それぞれ（ａ）の流量Ｑを０とする場合、（ｂ）の流量Ｑを１／２×設計流量Ｑ_dの場合、（ｃ）の流量Ｑを設計流量Ｑ_dの場合として示した。これら２図の対比からわかるように、本発明の遠心羽根車を装備した遠心ポンプは、（ａ）の流量Ｑを０とする場合、（ｂ）の流量Ｑを１／２×設計流量Ｑ_dの場合、（ｃ）の流量Ｑを設計流量Ｑ_dの場合のいずれにおいても圧力変動は小さくなり、ポンプの圧力脈動が大幅に低減されていることがわかる。
【００３０】
【発明の効果】
以上説明したように、従来の液体用の遠心ポンプと比べ、本発明は、羽根車出口における軸方向流路幅を羽根圧力面側と羽根負圧面側で相違するものとし、特に羽根圧力面側の軸方向流路幅に比べ羽根負圧面側の軸方向流路幅が拡大する、または、徐々に拡大してゆく、あるいは、羽根入口側における羽根圧力面から羽根負圧面までの羽根間で、羽根圧力面側の軸方向流路幅が羽根負圧面側の軸方向流路幅と同じで、かつ、羽根圧力面側の軸方向流路幅と羽根負圧面側の軸方向流路幅が同じである状態の羽根車入口から羽根車出口になるに連れて羽根圧力面側の軸方向流路幅に比して羽根負圧面側の軸方向流路幅が徐々に縮小して形成されていることで、液体用の遠心ポンプのイニシャルコストを格別にかけることなく、液体用の遠心ポンプの高効率化、さらに、遠心ポンプとした場合でのポンプの高キャビテーション性能を維持しながら、液体用の遠心ポンプの圧力脈動を低減化し、さらに騒音の低減化を図ることができ、本発明は従来にない優れた効果を奏するものである。
【図面の簡単な説明】
【図１】羽根車と渦巻室の断面図である。
【図２】従来の羽根車の構造を模式的に説明する断面図である。
【図３】本発明の羽根車出口における流れ構造および圧力脈動の説明図である。
【図４】本発明の液体用の遠心ポンプの一実施の形態の羽根車の構造を模式的に説明する断面図である。
【図５】本発明の液体用の遠心ポンプの他の実施の形態の羽根車の構造を模式的に説明する断面図である。
【図６】本発明の液体用の遠心ポンプの他の実施の形態の羽根車の構造を模式的に説明する断面図である。
【図７】本発明の液体用の遠心ポンプのさらに他の実施の形態の羽根車の構造を模式的に説明する断面図である。
【図８】従来の羽根車の遠心ポンプと本発明の羽根車を有する遠心ポンプの性能比較図である。
【図９】従来の羽根車と本発明の羽根車のＮＰＳＨ３（３％揚程低下）測定比較図である。
【図１０】従来の羽根車出口における圧力脈動波形図である。
【図１１】本発明の羽根車出口における圧力脈動波形図である。
【符号の説明】
１羽根車
２渦巻室
３舌部
４羽根車主板
５羽根
６羽根車側板
７羽根圧力面
８羽根負圧面
９吸込口
１０吐出口
１１羽根車外径
１２羽根間流路
１３吐出管
１４羽根車入口
１５羽根車出口
１６軸方向流路幅
１７軸方向流路幅
１８隙間
１９主流
２０低エネルギー流体[0001]
BACKGROUND OF THE INVENTION
The present invention relates to the structure of a centrifugal impeller of a centrifugal pump for liquid, among the structures of a centrifugal impeller included in a centrifugal fluid machine.
[0002]
[Prior art]
Conventionally, in the centrifugal impeller (hereinafter referred to as “impeller”) 1 of the liquid centrifugal pump shown in FIG. 1, the blades from the blade pressure surface 7 to the blade negative pressure surface 8 of the impeller, as shown in FIG. 2. 12 or in the range of Z × θ [θ = 360 ° / Z, Z: number of blades] between the blades from the blade pressure surface to the blade suction surface when the blade thickness is ignored, and the impeller At each blade radius from the inlet 14 toward the impeller outlet 15, the axial flow path width 16 is constant. By the way, the Coriolis force of 2ρωW per unit area acts on the fluid of density ρ flowing at the total speed W in the impeller 1 rotating at the angular velocity ω more than other forces. On the other hand, it is considered that the axial flow path width 16 is narrowed and the Coriolis force of 2ρωW is further increased with respect to the centrifugal pump for a small low specific speed liquid . Therefore, as shown in FIG. 3B, in the boundary layer near the surface of the impeller main plate 4 or the surface of the impeller side plate 6, the fluid flows from the blade pressure surface 7 between the blades 5 to the blade negative pressure surface 8. The low energy fluid 20 accumulates in the vicinity of the blade suction surface 8, particularly in the vicinity of the surface of the impeller side plate 6 and the corner of the blade suction surface 8, and the secondary flow that is the low energy fluid 20 is a flow passage. The effective flow path width of the non-low energy fluid which is the main flow 19 in the vicinity of the blade suction surface 8 is remarkably narrowed. For this reason, the pressure is not uniform between the blades 5, and in FIG. 1, the pressure at the discharge port 10 is caused by the potential interference between the flow at the impeller outlet 15 and the tongue 3 of the spiral chamber 2 and the viscous wake interference. , (Cycle number n) × (blade number z), and cyclic fluctuation ΔP (pressure pulsation) having a basic component is repeated.
[0003]
In order to reduce the vibration and noise of the centrifugal pump for liquid , conventionally, the gap 18 between the tongue 3 of the spiral chamber 2 and the outer diameter 11 of the impeller 1 shown in FIG. Measures such as changing the shape of 3 have been taken. By doing so, vibration and noise or pressure pulsation of the centrifugal pump for liquid can be reduced to some extent. However, there may be a lot of fluid loss between the impeller 1 and the spiral chamber 2, and the performance of the liquid centrifugal pump tends to decrease.
[0004]
On the other hand, when the impeller 1 of the centrifugal pump for liquid is the impeller 1 of a centrifugal pump (hereinafter referred to as “pump”), the number of impellers of the impeller 1 can be reduced or Cavitation performance of the pump can be improved by slowing the flow velocity at the vehicle entrance 14 or optimizing the shape of the blades 5. However, there is a trade-off relationship between pump cavitation performance and pump efficiency, which has a problem of adversely affecting the high efficiency of the pump.
[0005]
[Patent Document 1]
Japanese Patent Application No. 2002-119964 (filed on April 22, 2002)
[0006]
[Problems to be solved by the invention]
In the present invention, the initial cost without a centrifugal pump for liquid, while maintaining a high cavitation performance in the case of a high efficiency or pump of the centrifugal pump for liquids, the tongue of the impeller outlet flow and swirl chamber It is an object of the present invention to provide a liquid centrifugal pump capable of reducing pressure pulsation and noise by reducing potential interference and viscous wake interference.
[0007]
[Means for Solving the Problems]
The means of the present invention for solving the above-mentioned problem is that the fluid is sucked into the suction port 9 by being driven by a motor, and the spiral chamber around the impeller 1 is caused by centrifugal force from the impeller 1 in the spiral chamber 2. 2 is a liquid centrifugal pump comprising a discharge pipe 13 for discharging the discharged fluid from the peripheral portion of the spiral chamber 2 and discharging it to the discharge port 10 in a lump, and this liquid centrifugal pump The impeller 1 is the most important element. The impeller 1 is composed of a plurality of blades 5, an impeller main plate 4 and an impeller side plate 6, and the two adjacent blades 5 and blades of the impeller 1. The inter-blade channel 12 formed by the car main plate 4 and the impeller side plate 6 has an axial channel width 16 on the vane pressure surface 7 side as the vane negative pressure surface 8 while moving from the impeller inlet 14 to the impeller outlet 15. The centrifugal machine is different from the axial flow path width 17 on the side. Than is. In this case, between the blades 5 from the blade pressure surface 7 to the blade negative pressure surface 8 and from the impeller inlet 14 to the impeller outlet 15, the axial flow path width 16 on the blade pressure surface 7 side and the blade negative Those having different axial flow path widths 17 on the pressure surface 8 side are also included in the centrifugal granule machine.
[0008]
By doing so, during the operation of the liquid centrifugal pump , the flow of the impeller 1 in the inter-blade flow path 12 or the secondary flow of the low energy fluid 20 generated by the centrifugal force and the Coriolis force and the exit of the impeller It is possible to control 15 flow distributions.
[0009]
That is, in the first means , in the centrifugal pump for liquid having a centrifugal impeller, the impeller 1 is disposed between the blade pressure surface 7 and the blade negative pressure surface 8 on the impeller outlet 15 side, between the blade pressure surfaces 7 and the blade pressure surface. Compared to the axial flow path width 16 on the 7 side, the axial flow path width 17 on the blade suction surface 8 side is increased, and further between the blades 5 from the blade pressure surface 7 to the blade suction surface 8 on the blade inlet 14 side. The axial flow path width 17 on the blade negative pressure surface 8 side is larger than the axial flow path width 16 on the blade pressure surface 7 side, and compared with the axial flow path width 16 on the blade pressure surface 7 side. Thus , the liquid centrifugal pump is formed from the impeller inlet 14 to the impeller outlet 15 while maintaining the state in which the axial flow path width 17 on the impeller negative pressure surface 8 side is enlarged .
In the second means, in the centrifugal pump for liquid having a centrifugal impeller , the impeller 1 is disposed between the blade pressure surface 7 and the blade negative pressure surface 8 on the impeller outlet 15 side, between the blade pressure surface 7 side and the blade pressure surface 7 side. The axial flow passage width 17 on the blade suction surface 8 side is larger than the axial flow passage width 16 of the blade, and further between the blades 5 from the blade pressure surface 7 to the blade suction surface 8 on the blade inlet 14 side, the blade The axial flow path width 16 on the pressure surface 7 side is the same as the axial flow path width 17 on the blade negative pressure surface 8 side, and the axial flow path width 16 on the blade pressure surface 7 side and the shaft on the blade negative pressure surface 8 side. As the impeller inlet 14 in the same state as the directional flow path width 17 changes to the impeller outlet 15, the axial flow path width between the blade pressure surface 7 side and the blade negative pressure surface 8 side gradually decreases. The axial flow path width 16 on the blade pressure surface 7 side is formed smaller than the axial flow path width 17 on the blade negative pressure surface 8 side. It is a centrifugal pump for liquids, characterized in that.
[0010]
By doing so, the effective flow width of the main flow in the vicinity of the blade suction surface 8 is increased, and it is considered that the circumferential distribution of the flow at the impeller outlet 15 is made uniform to some extent. By reducing the interference with the tongue 3 of the spiral chamber 2, pressure pulsation and noise of the liquid centrifugal pump can be reduced.
[0011]
On the other hand, also in the impeller inlet 14, the axial flow path width 17 between the impeller main plate 4 and the impeller side plate 6 on the impeller negative pressure surface 8 side is the impeller main plate 4 and impeller side plate 6 on the impeller pressure surface 7 side. It is the same as the axial flow path width 16 between them, or is gradually enlarged as compared with the axial flow path width 16 . By doing in this way, the high cavitation performance of the pump when it is set as the impeller 1 of the centrifugal pump for liquids, for example, the impeller 1 of a centrifugal pump, can be maintained.
[0012]
In the third means , the impeller 1 is a small low specific speed whose outer diameter is 200 mm or less and whose specific speed Ns = nQ ^1/2 / H ^3/4 is 200 or less . A centrifugal pump for liquid of the first or second means .
[0013]
In the fourth means , the impeller 1 has an axial flow on the blade suction surface 8 side with respect to the axial flow path width 16 on the blade pressure surface 7 side between the two blades 5 and 5 facing each other at the maximum outer diameter. The centrifugal pump for liquid of any one of the first to third means, wherein the path width 17 is larger than 1 and 4 or less.
[0014]
By the way, generally considering the performance of the centrifugal pump for liquid, the axial flow path width at the impeller inlet 14 is made equal to or larger than the axial flow path width at the impeller outlet 15. Based on that. In addition, the ratio of the axial flow path width 17 on the blade negative pressure surface 8 side to the axial flow path width 16 on the blade pressure surface 7 side on the impeller inlet 14 side may be different. On the other hand, when only the reduction of pressure pulsation is emphasized, the axial flow passage width 17 on the blade negative pressure surface 8 side is set to the axial flow passage width 16 on the blade pressure surface 7 side on the impeller inlet 14 side as in the prior art. It may be the same.
[0015]
Here, the operation of the present invention will be described. First, in order to reduce the vibration and noise of a conventional centrifugal pump for liquid , the gap 18 between the tongue 3 of the spiral chamber 2 and the outer diameter 11 of the impeller is widened, or the shape of the tongue 3 of the spiral chamber 2 is changed. The vibration and noise of the centrifugal pump for liquid could be reduced to some extent by changing it. However, it may fluid loss between the impeller 1 and the swirl chamber 2 a number occurs, there is a tendency to decreased performance of the centrifugal pump for liquids, also the size up centrifugal pump for liquids There is also a tendency to try. On the other hand, regarding the cavitation performance of the pump in the case of a liquid centrifugal pump , the number of blades 5 of the impeller 1 is reduced, the flow velocity at the impeller inlet 14 is decreased, the shape of the impeller inlet 14 and the blade 5 The cavitation performance can be improved by reducing the number of blades, reducing the flow velocity at the impeller inlet 14, and optimizing the shape of the blades 5. However, there is a trade-off relationship between pump cavitation performance and pump efficiency, which adversely affects maintaining high efficiency of the pump. On the other hand, in the fluid machine of the present invention, by adopting the impeller 1 according to the means of claim 1, high efficiency of the centrifugal pump for liquid and high pump cavitation performance when the centrifugal pump is used are maintained. Pressure fluctuations can be greatly reduced.
[0016]
DETAILED DESCRIPTION OF THE INVENTION
Embodiments of the present invention will be described with reference to the drawings. FIG. 1 is a cross-sectional view of an impeller and a spiral chamber. FIG. 2 is a cross-sectional view schematically illustrating the structure of an impeller of a conventional centrifugal pump for liquid by linearly showing the end surfaces cut at the inner diameter side (a) and outer diameter side (b), and the impeller It is a schematic diagram which shows the part for radius. 3A and 3B are explanatory views of a conventional fluid flow structure and pressure pulsation at an impeller outlet, in which FIG. 3A is a schematic diagram of pressure pulsation, and FIG. 3B is a schematic diagram of a flow structure in the impeller. FIG. 4 illustrates the structure of an impeller according to an embodiment of the liquid centrifugal pump of the present invention by schematically showing linearly the end faces cut at the inner diameter side (a) and the outer diameter side (b). It is a schematic diagram which shows sectional drawing and the part for the radius of an impeller. FIG. 5 schematically shows a structure of an impeller according to another embodiment of the liquid centrifugal pump of the present invention, with end faces cut at the inner diameter side (a) and the outer diameter side (b) schematically shown as straight lines. It is a schematic diagram which shows sectional drawing to demonstrate and the part for the radius of an impeller. FIG. 6 schematically shows a structure of an impeller of still another embodiment of the centrifugal pump for liquid according to the present invention, with the end faces cut at the inner diameter side (a) and the outer diameter side (b) schematically linearized. FIG. 2 is a cross-sectional view described above and a schematic diagram showing the radius of an impeller. FIG. 7 schematically shows the structure of the impeller of still another embodiment of the centrifugal pump for liquid according to the present invention, with the end faces cut at the inner diameter side (a) and the outer diameter side (b) schematically linearized. FIG. 2 is a cross-sectional view described above and a schematic diagram showing the radius of an impeller. FIG. 8 is a performance comparison diagram of a conventional impeller centrifugal pump and an impeller centrifugal pump of the present invention. FIG. 9 is a measurement comparison diagram of NPSH3 (3% reduction in lift) of a conventional impeller centrifugal pump and the impeller of the present invention. FIG. 10 is a diagram showing a pressure pulsation waveform of a conventional impeller. FIG. 11 is a diagram showing a pressure pulsation waveform of the impeller of the present invention.
[0017]
One embodiment of the present invention is a centrifugal pump for liquid . As shown in FIG. 1, this centrifugal pump has an impeller 1 for sucking a liquid pumped up into a spiral chamber 2 by driving a motor into a suction port 9 of the centrifugal pump and discharging it to a discharge port 10 of the centrifugal pump. And a spiral chamber 2 that collects liquid discharged from the impeller 1. The impeller 1 is the most important element in the centrifugal pump for liquid according to the present invention. The impeller 1 includes an impeller main plate 4 and an impeller side plate 6 that vertically sandwich and hold a plurality of blades 5 and the blades 5. Is formed. As shown in FIG. 4, the inter-blade flow path 12 formed by the two adjacent blades 5 of the impeller 1 and the impeller main plate 4 and the impeller side plate 6 has an axial flow passage width on the blade negative pressure surface 8 side. 17 is formed from the impeller inlet 14 to the impeller outlet 15 while maintaining a state where 17 is gradually enlarged compared to the axial flow path width 16 on the blade pressure surface 7 side.
[0018]
By doing in this way, the flow effective flow path near the blade negative pressure surface 8 of the impeller 1 is expanded, and the flow of the flow path 12 between the blades including the flow on the impeller outlet 15 side is made uniform to some extent. By reducing the interference between the flow of the vehicle outlet 15 and the tongue 3 of the spiral chamber 2, pressure pulsation and noise of the centrifugal pump for liquid can be reduced. On the other hand, at the impeller inlet 14 of the impeller 1, the axial flow path width 17 on the blade negative pressure surface 8 side is gradually enlarged compared to the axial flow path width 16 on the blade pressure surface 7 side, and the impeller inlet 14 side Since the relative flow velocity at the nearby blade suction surface 8 is also reduced, the high cavitation performance of the pump that is the impeller 1 of the centrifugal pump can be maintained.
[0019]
Furthermore, in the embodiment shown in FIG. 5, in the portion on the inner diameter side of the impeller 1 on the impeller inlet 14 side as shown in FIG. 5A, the axial flow passage width is negative on the impeller negative pressure surface 7 side. Both pressure sides 8 also have the same axial flow path width. However, as shown in FIG. 5 (b), at the outer diameter side portion of the impeller 1 on the impeller outlet 15 side, the blade negative pressure surface 8 is compared with the axial flow path width 16 on the blade pressure surface 7 side. The axial flow path width 17 on the side is gradually enlarged.
[0020]
Further, in the embodiment shown in FIG. 6, as in the embodiment in FIG. 5, as shown in FIG. 5A, in the portion on the inner diameter side of the impeller 1 that is on the impeller inlet 14 side, The passage width is the same axial passage width on the blade pressure surface 7 side and on the blade negative pressure surface 8 side. However, as shown in FIG. 6B, the impeller main plate 4 has a constant thickness at the outer diameter side portion of the impeller 1 on the impeller exit 15 side, but the impeller side plate 6 has a constant thickness. The plate thickness is tapered only on the inner surface side, so that the axial flow passage width 17 on the blade negative pressure surface 8 side is gradually enlarged compared to the axial flow passage width 16 on the blade pressure surface 7 side. Yes.
[0021]
Further, in the embodiment shown in FIG. 7, as shown in FIG. 7A, in the portion on the inner diameter side of the impeller 1 on the impeller inlet 14 side, the axial flow path width is the blade negative pressure surface on the blade pressure surface 7 side. Both 8 sides have the same axial flow path width. However, as shown in FIG. 7B, in the outer diameter side portion of the impeller 1 on the impeller outlet 15 side, the axial flow path width 16 on the impeller pressure surface 7 side is directed to the impeller negative pressure surface 8 side. The axial flow path width 16 has a constant narrow width for a certain length, but the axial flow path width 17 on the blade negative pressure surface 8 side increases from the middle toward the blade negative pressure surface 8 side. It will gradually expand.
[0022]
As described above, the centrifugal pump which is an embodiment of the liquid centrifugal pump of the present invention can achieve low pressure pulsation and low noise.
[0023]
Further, in another embodiment, the small low ratio in which the axial flow passage width 17 on the blade negative pressure surface 8 side with respect to the axial flow passage width 16 on the blade pressure surface 7 side is greater than 1 and 4 or less. Speed centrifugal pump.
[0024]
【Example】
First, the dimensions of the specifications of the centrifugal pump of the embodiment of the present invention shown in FIG. 4 are exemplified. The outer diameter of the impeller 1 is 62 mm, and the axial flow path width of the blade suction surface 8 of the impeller inlet 14 is 5. 8 mm, the axial flow width of the blade pressure surface 7 of the impeller inlet 14 is 3.8 mm, the axial flow width of the blade negative pressure surface 8 of the impeller outlet 15 is 4.5 mm, and the blade pressure surface of the impeller outlet 15 The axial flow path width of 7 is 2.5 mm.
[0025]
Furthermore, the dimensions of the specifications of the centrifugal pump of the embodiment of the present invention shown in FIG. 5 are exemplified. The outer diameter of the impeller 1 is 62 mm, the axial flow path width of the impeller inlet 14 is the blade negative pressure surface 8 and the blade pressure. The surface 7 has the same width of 4.8 mm, the axial flow width of the blade suction surface 8 of the impeller outlet 15 is 4.5 mm, and the axial flow width of the blade pressure surface 7 of the impeller outlet 15 is 2.5 mm. is there.
[0026]
For comparison, when the dimensions of a centrifugal pump in which the conventional axial flow path width is the same on both the blade suction surface side and the blade pressure surface side are shown, the outer diameter of the impeller 1 is 62 mm, the impeller inlet 14 is The axial flow path width is 4.8 mm, and the axial flow path width of the impeller outlet 15 is 3.5 mm. In the above, all other dimensions of the centrifugal pump including the spiral chamber 2 are the same, and the design flow rate Q _d is 30 l / min.
[0027]
FIG. 8 compares the performance of a centrifugal pump equipped with a conventional impeller and a centrifugal pump equipped with the impeller 1 of the embodiment shown in FIG. 4 in the present invention. However, the performance of the centrifugal pump equipped with the impeller 1 of the embodiment shown in FIG. 5 of the present invention matches the performance of the centrifugal pump equipped with the impeller 1 of the embodiment shown in FIG. The performance of the centrifugal pump equipped with the impeller 1 shown in FIG. 4 is compared with the centrifugal pump equipped with the conventional impeller. In the entire flow rate range, the centrifugal pump efficiency equipped with the impeller in the present invention is not different from the efficiency of the conventional pump, but the total lift H appears a little higher, and the internal flow and outlet flow of the centrifugal impeller are controlled. I guess that.
[0028]
FIG. 9 shows a comparison of the 3% reduction in the head of the small low specific speed centrifugal pump equipped with the conventional impeller and the small low specific speed centrifugal pump equipped with the impeller 1 of the embodiment of FIG. 4 in the present invention. Show. There is a tendency that the cavitation performance of the centrifugal pump having the centrifugal impeller of the present invention is slightly improved as the pumped water amount Q increases as compared with the conventional pump.
[0029]
FIG. 10 shows the pressure fluctuation at the discharge port of the centrifugal pump equipped with the conventional impeller, and FIG. 11 shows the pressure fluctuation at the discharge port of the centrifugal pump equipped with the impeller of the embodiment shown in FIG. When the flow rate Q in (a) is 0, the flow rate Q in (b) is 1/2 × design flow rate Q _d , and the flow rate Q in (c) is shown as design flow rate Q _d . As can be seen from the comparison of these two figures, the centrifugal pump equipped with the centrifugal impeller of the present invention has the flow rate Q of (b) set to 1/2 × design flow rate Q _d when the flow rate Q of (a) is set to 0. cases, it can be seen that even pressure variation in any of the cases of the flow rate Q of the design flow rate Q _d is reduced, the pressure pulsation of the pump is significantly reduced in (c).
[0030]
【The invention's effect】
As described above, the present invention differs from the conventional centrifugal pump for liquid in that the axial flow path width at the impeller outlet is different between the blade pressure surface side and the blade negative pressure surface side. The axial flow path width on the blade suction surface side expands or gradually expands compared to the axial flow path width of, or between the blades from the blade pressure surface to the blade suction surface on the blade inlet side, The axial flow path width on the blade pressure surface side is the same as the axial flow path width on the blade suction surface side, and the axial flow path width on the blade pressure surface side is the same as the axial flow path width on the blade suction surface side. The axial flow passage width on the blade negative pressure surface side is gradually reduced as compared to the axial flow passage width on the blade pressure surface side as the impeller inlet in the state is changed to the impeller exit. it is, particularly without imposing initial cost of the centrifugal pump for liquid, a centrifugal pump for liquids High efficiency, further, while maintaining a high cavitation performance of the pump in the case of a centrifugal pump, to reduce the pressure pulsation of a centrifugal pump for liquid, it is possible to further achieve a reduction of noise, the present invention is conventional It has an outstanding effect.
[Brief description of the drawings]
FIG. 1 is a cross-sectional view of an impeller and a spiral chamber.
FIG. 2 is a cross-sectional view schematically illustrating the structure of a conventional impeller.
FIG. 3 is an explanatory diagram of the flow structure and pressure pulsation at the impeller outlet of the present invention.
FIG. 4 is a cross-sectional view schematically illustrating the structure of an impeller according to an embodiment of the liquid centrifugal pump of the present invention.
FIG. 5 is a cross-sectional view schematically illustrating the structure of an impeller according to another embodiment of the liquid centrifugal pump of the present invention.
FIG. 6 is a cross-sectional view schematically illustrating the structure of an impeller according to another embodiment of the liquid centrifugal pump of the present invention.
FIG. 7 is a cross-sectional view schematically illustrating the structure of an impeller according to still another embodiment of the liquid centrifugal pump of the present invention.
FIG. 8 is a performance comparison diagram of a centrifugal pump of a conventional impeller and a centrifugal pump having the impeller of the present invention.
FIG. 9 is a NPSH3 (3% reduction in lift) measurement comparison diagram of a conventional impeller and the impeller of the present invention.
FIG. 10 is a pressure pulsation waveform diagram at a conventional impeller outlet.
FIG. 11 is a pressure pulsation waveform diagram at the impeller outlet of the present invention.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 Impeller 2 Swirl chamber 3 Tongue part 4 Impeller main plate 5 Blade 6 Impeller side plate 7 Blade pressure surface 8 Blade negative pressure surface 9 Suction port 10 Discharge port 11 Impeller outer diameter 12 Flow channel between blades 13 Discharge pipe 14 Impeller inlet 15 Impeller outlet 16 Axial channel width 17 Axial channel width 18 Gap 19 Main flow 20 Low energy fluid

Claims

In a centrifugal pump for a liquid having a centrifugal impeller, the centrifugal impeller includes a blade between the blade pressure surface and the blade negative pressure surface on the impeller outlet side compared to the axial flow path width on the blade pressure surface side. The axial flow path width on the suction surface side is increased, and further, between the blades from the blade pressure surface on the blade inlet side to the blade suction surface, the axial flow width on the blade pressure surface side is larger than the axial flow path width on the blade pressure surface side. From the impeller inlet to the impeller outlet while the axial passage width is enlarged and the axial passage width on the vane suction surface side is enlarged compared to the axial passage width on the vane pressure surface side A centrifugal pump for liquid , characterized in that it is formed .

In a centrifugal pump for a liquid having a centrifugal impeller, the centrifugal impeller includes a blade between the blade pressure surface and the blade negative pressure surface on the impeller outlet side compared to the axial flow path width on the blade pressure surface side. The axial flow path width on the suction side increases, and the axial flow path width on the blade pressure side is between the blades from the blade pressure surface on the blade inlet side to the blade suction surface. The blade pressure surface is changed from the impeller inlet to the impeller outlet in the same state as the passage width and the same in the axial passage width on the blade pressure surface side and the axial passage width on the blade suction surface side. The axial flow path width between the side and the blade suction surface side is gradually reduced, and the axial flow path width on the blade pressure surface side is formed to be smaller than the axial flow path width on the blade suction surface side. A centrifugal pump for liquids characterized by the above.

The centrifugal impeller has a small low specific speed with an outer diameter of 200 mm or less and a specific speed Ns = nQ ^1/2 / H ^3/4 of 200 or less. Centrifugal pump for liquids .

In the centrifugal impeller, the axial flow path width on the blade suction surface side with respect to the axial flow path width on the blade pressure surface side is greater than 1 and less than or equal to 4 between the two facing blades at the maximum outer diameter The centrifugal pump for liquid according to claim 1, wherein the centrifugal pump is for liquid .