JPS61190184A - Screw rotor of screw compressor - Google Patents
Screw rotor of screw compressorInfo
- Publication number
- JPS61190184A JPS61190184A JP3082185A JP3082185A JPS61190184A JP S61190184 A JPS61190184 A JP S61190184A JP 3082185 A JP3082185 A JP 3082185A JP 3082185 A JP3082185 A JP 3082185A JP S61190184 A JPS61190184 A JP S61190184A
- Authority
- JP
- Japan
- Prior art keywords
- rotors
- tooth
- rotor
- gap
- tooth profile
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Granted
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
- F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
- F04C18/082—Details specially related to intermeshing engagement type pumps
- F04C18/084—Toothed wheels
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
- F04C18/08—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
- F04C18/12—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
- F04C18/14—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
- F04C18/16—Rotary-piston pumps specially adapted for elastic fluids of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Applications Or Details Of Rotary Compressors (AREA)
Abstract
Description
【発明の詳細な説明】
弦術分団
本発明は、スクリュー圧縮機のスクリューロータに関し
、詳しくは、雌雄両ロータをタイミングギ′ヤ等の同期
装置により両ロータを互いに接触しないように同期回転
せしめるタイプの無潤滑式スクリュー圧縮機におけるロ
ータに関する。[Detailed Description of the Invention] The present invention relates to a screw rotor for a screw compressor, and more specifically to a type in which male and female rotors are rotated synchronously using a synchronizing device such as a timing gear so that the two rotors do not come into contact with each other. This invention relates to a rotor in a non-lubricated screw compressor.
従米技暫
上記タイプの圧縮機は、油冷却タイプの圧縮機に較べる
と、ロータは昇温し易＜ ２０ （１”Ｃ以上にもなる
。従って、このタイプの圧縮機の設計においては、両ロ
ータの熱膨張を十分考慮して運転中に両ロータが干渉す
ることのないようにしなければならない。Compared to oil-cooled compressors, the rotor of the above type of compressor is more likely to heat up to < 20 (1"C or more). Therefore, in the design of this type of compressor, both The thermal expansion of the rotors must be taken into consideration to prevent interference between the two rotors during operation.
従来は、運転中の両ロータの熱膨張による干渉を回避す
るため、両ロータの軸間距離（ＣＩ））を十分数るよう
にしている。すなわち、今、両ロータの線膨張係数をβ
、設定上昇温度をＴとして膨張量Δβを次式：
％式％
で求め、雄ロータの歯先円と雌ロータの歯底円間の離隔
寸法（歯面間隙間）が、上記膨張量に、運転中最低必要
とされる安全隙間を加えた値になるように、両ロータの
細心位置を決定している。Conventionally, in order to avoid interference due to thermal expansion of both rotors during operation, the distance between the axes (CI) of both rotors is set to be a sufficient number. In other words, now the coefficient of linear expansion of both rotors is β
, the expansion amount Δβ is determined by the following formula, with the set temperature increase being T: % Formula % The distance between the tip circle of the male rotor and the root circle of the female rotor (gap between tooth flanks) is equal to the above expansion amount, The positions of both rotors are carefully determined so that the minimum safety gap required during operation is added.
ところで、圧縮効率を大ならしめるためには、両ロータ
間のシールラインにおける歯面間隙間を最小として圧縮
空気の吸気側への漏れを少なくすることが重要であるが
、この歯面間隙間は、両ロータの軸心位置のみならず両
ロータの膨張した状態の歯形により決定される。By the way, in order to increase the compression efficiency, it is important to minimize the gap between the tooth surfaces at the seal line between both rotors to reduce the leakage of compressed air to the intake side. , is determined not only by the axial center positions of both rotors but also by the tooth profiles of both rotors in the expanded state.
ところが、前記したように、従来は、歯面間隙間を最小
ならしめるため、両ロータの膨張歯形を考慮することは
なかった。したがって、従来は常温時におけるロータ歯
形をいくら理想形状としても、運転時すなわち昇温時に
おいては圧縮効率が悪化せざるを得なかった。However, as described above, in the past, in order to minimize the gap between the tooth surfaces, the expansion tooth profiles of both rotors were not considered. Therefore, conventionally, no matter how ideal the rotor tooth profile is at room temperature, the compression efficiency inevitably deteriorates during operation, that is, when the temperature rises.
因に、第３図（Ｉ）、（ＩＩ）に示すように、雌雄ロー
タＦ、Ｍは両者共に、常温時歯形（実線で示す）と昇温
時歯形（一点鎖線で示す膨張歯形）とは大巾に異る。そ
して、図に明らかなように、熱膨張は均等に生ずるもの
ではなく、大概、各歯形点か軸心に近い程、該歯形点の
膨張量は少いと云えるが、厳密には、有限要素法等の数
学的手法により各歯形点の膨張量を求めることができる
。Incidentally, as shown in FIGS. 3(I) and (II), for both the male and female rotors F and M, the tooth profile at room temperature (shown by the solid line) and the tooth profile at elevated temperature (the expanded tooth profile shown by the dashed-dotted line) are different. Very different. As is clear from the figure, thermal expansion does not occur uniformly, and it can be said that the closer each tooth profile point is to the axis, the smaller the amount of expansion at that tooth profile point. The amount of expansion at each tooth profile point can be determined using a mathematical method such as the method.
本発明０技物的Ｗ顆
したがって、本発明の解決すべべ技術的課題は、無潤滑
式スクリュー圧縮磯において、運転時における雌雄両ロ
ータの熱膨張歯形を考慮して常温時における両ロータの
歯形を決定することにより圧縮機の圧縮効率を向−１ニ
させることに存する。Therefore, the technical problem to be solved by the present invention is to solve the problem of the tooth profile of both male and female rotors at room temperature by considering the thermal expansion tooth profile of both the male and female rotors during operation in a non-lubricated screw compression rock. The purpose is to improve the compression efficiency of the compressor by determining the .
本発明外に旨
本発明は、雌雄各ロータ表面に焼付防市川コーティング
被膜を塗布してなるスクリュー圧縮機のスクリューロー
タに適用される。In addition to the present invention, the present invention is applied to a screw rotor of a screw compressor in which an anti-seizure Ichikawa coating film is applied to the surfaces of each of the male and female rotors.
本発明に係るスクリューロータは、常温において歯面間
隙間が［（）」となるように配置した雌雄ロータを想定
するとともに、運転時に生ずる熱膨張に起因する上記両
ロータの軸直角断面歯形における各歯形点の干渉量を算
出し、該各歯形点の干渉量に、運転時に最低限与えるベ
ト安全隙間を加算した値を常温時の両ロータ間における
各歯形点間の歯面間隙間として保持するように両ロータ
歯形を決定してなることを特徴とし、かつ上記歯面間隙
間が上記両ロータの焼付防止コーティング被膜の総輪断
面膜厚より小さい歯形点については、上記歯面間隙間を
上記焼付防止コーティング被膜の総軸断面膜厚と等しい
値にすることを特徴としている。The screw rotor according to the present invention assumes male and female rotors arranged so that the gap between the tooth surfaces is [()'' at room temperature, and the tooth profile of the above-mentioned rotors perpendicular to the axis due to thermal expansion that occurs during operation. The amount of interference between the tooth profile points is calculated, and the sum of the amount of interference at each tooth profile point plus the minimum sticky safety clearance provided during operation is held as the gap between the tooth surfaces between each tooth profile point between both rotors at room temperature. The tooth profile of both rotors is determined as follows, and for tooth profile points where the gap between the tooth flanks is smaller than the total cross-sectional film thickness of the anti-seize coating of both rotors, the gap between the tooth flanks is determined as above. It is characterized by having a value equal to the total axial cross-sectional film thickness of the anti-seize coating film.
上記構成によれば、常温時のロータ歯形を、・運転時の
熱膨張を考慮して決定しているので、運転時には、両ロ
ータ間のシールラインにおける歯面間隙間は安全隙間す
なわち許容される最低寸法となり、そのため理想的なシ
ールラインを形成することができ、以って圧縮効率の向
上を図ることができる。According to the above configuration, the rotor tooth profile at room temperature is determined by considering thermal expansion during operation, so during operation, the gap between the tooth surfaces at the seal line between both rotors is a safety gap, that is, an allowable gap. This is the minimum dimension, and therefore an ideal seal line can be formed, thereby improving compression efficiency.
また、上記構成では、設定すべき歯面間隙間が両ロータ
に塗布される焼付防止コーティング被膜の総輪断面膜厚
より小さい歯形点については、上記歯面間隙間を上記焼
付防止コーティング被膜の総輪断面膜厚と等しい値にし
ているので、両口夕の圧縮機ハウノングに対する組（＝
Ｉけの際にも、両ロータのコーティング被膜を干渉させ
ることなく組付けられる。In addition, in the above configuration, for tooth profile points where the gap between the tooth flanks to be set is smaller than the total cross-sectional film thickness of the anti-seize coating coated on both rotors, the gap between the tooth flanks is Since the value is set equal to the ring cross-section film thickness, the set (=
Even when the rotors are installed, they can be assembled without interfering with the coatings on both rotors.
罰例
以下に、第１，２図に示した本発明の１実施例について
詳細に説明する。Penalty Example Below, one embodiment of the present invention shown in FIGS. 1 and 2 will be described in detail.
第１図（Ｉ）に、無潤滑式スクリュー圧縮磯の要部軸断
面を示している。この圧縮機に使用される雌雄両ロータ
Ｆ、Ｍは一般に非対称型と呼ばれる歯形を有する。図中
、実線は、常温時（２０℃）に雄ロータＭの歯先円Ｃｍ
と雌ロータＦの歯底円Ｂｆとの間の寸法すなわち歯面間
隙間が「０］になるように配置した各ロータの歯形を示
し、一方、図中破線は、圧縮機の運転中に生ずる各ロー
タの熱膨張、および運転中に雄ロータの外周円と雌ロー
タの歯底円との間に最低与えるべき安全隙間を考慮した
上で決定した雌ロータの歯形を示している。FIG. 1 (I) shows an axial cross-section of the main part of the non-lubricated screw compression rock. Both male and female rotors F and M used in this compressor have tooth profiles that are generally referred to as asymmetric types. In the figure, the solid line indicates the tip circle Cm of the male rotor M at room temperature (20°C).
The tooth profile of each rotor is shown so that the dimension between and the tooth bottom circle Bf of the female rotor F, that is, the gap between the tooth surfaces, is "0". The tooth profile of the female rotor is determined in consideration of the thermal expansion of each rotor and the minimum safety gap that should be provided between the outer circumferential circle of the male rotor and the root circle of the female rotor during operation.
第１図中実線で示された常温時の各ロータ歯形は次のと
おりである。The rotor tooth profiles at room temperature indicated by solid lines in FIG. 1 are as follows.
舞旦二名ｍ形
雌ロータＦは、その各歯のピッチ円Ｐｆの外側にアデン
ダムＡｆを有しかつ下記のロータ歯形を有する。The double m-shaped female rotor F has an addendum Af on the outside of the pitch circle Pf of each tooth, and has the following rotor tooth profile.
（前進歯形） ピッチ円（Ｐｆ）と、雌雄ロータの各中心点（Ｏｆ。(forward tooth profile) The pitch circle (Pf) and each center point (Of) of the male and female rotors.
Ｏｍ）を結ぶ線との交点（０１）を中心とする半径（Ｓ
Ｒ，）の円弧（ｄ２−８２）と、
半径（ｏｌｅ２）の延長線上に中心点（０２）を有する
半径（ＳＲ２）の円弧（ｅ２−ｆ２ ）と、半径（０２
−ｆ２）の上の点（Ｏ９）を中心とする半径ＳＲ３の円
弧（ｆ２ｇ２）と、
歯先円（Ｃｆ）上の円弧（ｇ２ａ２）と、を順に接続し
てなる。The radius (S) centered at the intersection (01) with the line connecting Om
An arc (d2-82) of radius (R,), an arc (e2-f2) of radius (SR2) having a center point (02) on an extension of radius (ole2), and an arc (e2-f2) of radius (02)
-f2) and a circular arc (f2g2) of radius SR3 centered on the point (O9) above the tooth tip circle (Cf) and an arc (g2a2) on the tip circle (Cf) are connected in order.
但し、点（ｄ２）は中心点（Ｏｆ、 Ｏｍ）を結ぶ線上
の点でかっ歯底円（Ｂｆ）上の点である。However, point (d2) is a point on the line connecting the center points (Of, Om) and is a point on the parenthesis circle (Bf).
点（［２）はピッチ円（Ｐｆ）より内側に位置している
。Point ([2) is located inside the pitch circle (Pf).
（追従側歯形）
雄ロータ（Ｍ）の円弧（ｄ、 −ｅｌ ）によって創成
される創成曲線（ｄ２−ｃ２）と、
７一
点（０５）を中心とする半径（ＳＲ５）の円弧（ｃ２−
１＋２）と、
ピッチ円（Ｐｆ）Ｊ：の点（０７）を中心とする半径（
ＳＲ？）の円弧（１１２−ａ２）と、を順に接続してな
る。(Following side tooth profile) A generation curve (d2-c2) created by the circular arc (d, -el) of the male rotor (M), and a circular arc (c2-c2) of radius (SR5) centered on point 7 (05)
1+2) and the radius centered at point (07) of the pitch circle (Pf) J: (
SR? ) are connected in order.
但し、点（ｌ］２）はピッチ円（Ｐｆ）上の点である。However, the point (l]2) is a point on the pitch circle (Pf).
雄ロータ歯形
雄ロータＭは、その各歯底のピッチ円Ｐｍの内側に上記
アデンダムＡｆに対応するデデンダムＤｍを有しかつ下
記のロータ歯形を有する。Male rotor tooth profile The male rotor M has a dedendum Dm corresponding to the addendum Af above on the inside of the pitch circle Pm at the bottom of each tooth, and has the following rotor tooth profile.
（前進側歯形） ピッチ円（Ｐｍ）と、雌雄ロータの各中心点（ｏｒ。(Advance tooth profile) The pitch circle (Pm) and each center point (or) of the male and female rotors.
Ｏｍ）を結ぶ点との交点（【０）を中心としかつ」二記
半径（ＳＲ，）に等しい半径の円弧（ｄ、 −ｅ、 ）
と、雌ロータ（Ｆ）の上記円弧（ｅ２ ｒ２）によっ
て創成される創成曲線（ｅ、 −ｆ、 ）と、雌ロータ
（Ｆ）の円’Ａ（ｆ２−ｇ２）によって創成される創成
曲線（Ｌｇ＋）と、
歯底円（Ｂｉ上の円弧（ｇ＋ａ＋）と、を順に接続して
なる。An arc (d, -e, ) whose center is the intersection point (0) with the point connecting Om) and whose radius is equal to the radius (SR, )
, the generated curve (e, −f, ) created by the above circular arc (e2 r2) of the female rotor (F), and the generated curve ( Lg+) and the root circle (arc (g+a+) on Bi) are connected in order.
一費一
但し、点（ｄｌ）は中心点（Ｏｆ、 Ｏｍ）を結ぶ線上
の点でかつ歯底円（Ｂ「）上の点である。However, the point (dl) is a point on the line connecting the center points (Of, Om) and a point on the root circle (B'').
点（ｆｌ）はピッチ円（Ｐｍ）より外方に位置している
。The point (fl) is located outward from the pitch circle (Pm).
（追従側歯形）
中心点（Ｏｆ、 Ｏｍ）を結ぶ線上の点（Ｏｌ）を中心
とする半径（ＳＲ，）の円弧（ｄ、 −ｃ、 ）と、雌
ロータ（Ｆ）の上記円弧（ｂ２−ｃ２）により創成され
る創成曲線（ｃｌ−ｂｌ）と、
ピッチ円（Ｐｎ）上の点（０８）を中心とする半径（Ｓ
Ｒ，）の円弧（ｂｌ−ａｌ）と、
を順に接続している。(Following side tooth profile) An arc (d, -c, ) of radius (SR,) centered on the point (Ol) on the line connecting the center points (Of, Om) and the above arc (b2) of the female rotor (F) -c2) and the radius (S) centered on the point (08) on the pitch circle (Pn).
The arc (bl-al) of R,) is connected in order.
但し、点（ｂ、）はピッチ円上の点である。However, point (b,) is a point on the pitch circle.
雌雄各ロータの歯形は上記のとおりであるが、この実施
例においては、さらに、両ロータ細心間距離は６７ＩＩ
Ｉ＋６、雄ロータＭの歯数は５枚、その歯先円Ｃｍの径
寸法は９０ＩＩＩＩ１１、そのピッチ円Ｐｍの径寸法は
６０，９０９ｍ、雌ロータＦの歯数は６枚、その歯先円
Ｃｆの径寸法は７６ＩＩＩＩａ、そのピッチ円Ｐｆの径
寸法は７３．０９１に夫々設定している。The tooth profiles of the male and female rotors are as described above, but in this example, the distance between the two rotors' centers is 67II.
I+6, the number of teeth of the male rotor M is 5, the diameter of its tip circle Cm is 90III11, the diameter of its pitch circle Pm is 60,909 m, the number of teeth of the female rotor F is 6, its tip circle Cf The diameter of the pitch circle Pf is set to 76IIIa, and the diameter of the pitch circle Pf is set to 73.091.
さて次に、圧縮機の運転時における各ロータの熱膨張を
考慮して常温時の各ロータ歯形を決定するため、第２図
（Ｉ−Ｖｌ）に示すように、第１図に示した常温時のロ
ータＭ、Ｆが熱膨張した歯形を算出し、該膨張歯形の仮
想ロータを回転させて各歯形点の干渉量を算出する。尚
、上記膨張歯形は各ロータが２００℃になった場合を想
定している。Next, in order to determine the tooth profile of each rotor at normal temperature in consideration of the thermal expansion of each rotor during operation of the compressor, as shown in Fig. 2 (I-Vl), The tooth profile of the rotors M and F at the time of thermal expansion is calculated, and the virtual rotor of the expanded tooth profile is rotated to calculate the amount of interference at each tooth profile point. Note that the expansion tooth profile described above is based on the assumption that each rotor is heated to 200°C.
〈第２図（■）〉
第２図（Ｉ）は第１図と同一のロータ噛み合い状態を示
しており、今この状態を各ロータ回転角ｒＯＪとする。<Fig. 2 (■)> Fig. 2 (I) shows the same rotor meshing state as in Fig. 1, and this state is now assumed to be each rotor rotation angle rOJ.
この回転状態では点ｄ１とｄ２とが接触してシールライ
ンの一点を形成する。点ｄ１〜ｄ２間の干渉量ε１゛は
１６０μｍとなる。In this rotating state, points d1 and d2 come into contact and form one point on the seal line. The amount of interference ε1' between points d1 and d2 is 160 μm.
〈第２図（ＩＩ）＞
第２図（１１）は、各ロータ回転角カ弓２°である場合
のロータ噛み合い状態を示している。この状態では、点
１〕１ とｂ２とが接触してシールラインの一点を形
成する。点ｂｌ−ｂ２開の干渉量ε２゛は６０μｍとな
る。<Figure 2 (II)> Figure 2 (11) shows the rotor meshing state when each rotor rotation angle is 2 degrees. In this state, point 1]1 and b2 contact to form one point on the seal line. The amount of interference ε2' at the point bl-b2 open is 60 μm.
〈第２図（ＩＩＩ）＞
第２図（Ｈｌ）は、各ロータ回転角が２４°である場合
のロータ噛み合い状態を示している。この状態では点ａ
、とａ２とが叉点ｇｌ とｇ２とが夫々接触して夫々シ
ールラインの一点を形成する。<Figure 2 (III)> Figure 2 (Hl) shows the rotor meshing state when each rotor rotation angle is 24 degrees. In this state, point a
, and a2 come into contact with each other at points gl and g2, respectively, to form one point of the seal line.
点ａ、 と８２間の干渉量ε３゛および点ｇ１ とｇ
２開の干渉量ε４゛は夫々１６０μｍとなる。The amount of interference ε3゛ between points a, and 82, and the points g1 and g
The amount of interference ε4' for the two openings is 160 μm.
〈第２図（ＩＶ）＞
第２図（ｎ）は、各ロータ回転角が４８°である場合の
ロータ噛み合い状態を示している。この状態では点ｄ、
と０２とが接触してシールラインの一点を形成する。点
ｄ１ と０２間の干渉量ε５゛は１３０μｍとなる。<Figure 2 (IV)> Figure 2 (n) shows the rotor meshing state when each rotor rotation angle is 48 degrees. In this state, point d,
and 02 come into contact to form one point of the seal line. The amount of interference ε5' between points d1 and 02 is 130 μm.
〈第２図（■）〉
第２図（Ｖ）は、各ロータ回転角が５４°である場合の
ロータ噛み合い状態を示している。この状態では点ｆ、
とｒ２とが接触してシールラインの一点を形成する
。点ｆ、と１２間の干渉量ε６゛は６０μｍとなる。<Figure 2 (■)> Figure 2 (V) shows the rotor meshing state when each rotor rotation angle is 54 degrees. In this state, point f,
and r2 contact to form one point of the seal line. The amount of interference ε6' between points f and 12 is 60 μm.
〈第２図（ｖｒ）＞
第２図（ｖｒ）は、各ロータ回転角が６６°である場合
のロータ噛み合い状態を示している。この状態では点ｅ
１ とｅ２ とが接触してシールラインの一点を形成
する。点ｅ１ と０２間の干渉量ε７゛は６０μｍとな
る。<Fig. 2 (vr)> Fig. 2 (vr) shows the rotor meshing state when each rotor rotation angle is 66°. In this state, point e
1 and e2 contact to form one point of the seal line. The amount of interference ε7' between points e1 and 02 is 60 μm.
上記の如くして両ロータの歯形間の熱膨張による干渉量
ε、゛〜ε７゛が求まれば、この干渉量ε１゛〜ε７゛
に基づいて、該干渉量ε１゛〜ε７゛を吸収し得るよう
に第１図中実線で示した常温時の各ロータ歯形を修正す
ればよい。この歯形修正は、雌雄ロータの何れか一方の
みについて行うことも、あるいは両ロータを共に行うこ
ともでとるが、本実施例では、雄ロータＭの歯形を基本
歯形としてこの修正は行なわず、雌ロータＦの歯形のみ
を修正・ している。すなわち、本実施例は、雌雄
両ロータの熱膨張による干渉を雌ロータの歯形修正によ
り吸収するようにしている。If the amount of interference ε, ``~ε7'' due to thermal expansion between the tooth profiles of both rotors is determined as described above, then the amount of interference ε1''~ε7'' is absorbed based on the amount of interference ε1~ε7''. To achieve this, the rotor tooth profile at room temperature shown by the solid line in FIG. 1 may be corrected. This tooth profile modification can be performed on either the male and female rotors, or both rotors, but in this embodiment, the tooth profile of the male rotor M is used as the basic tooth profile, and this modification is not performed on the female rotor. Only the tooth profile of rotor F has been corrected. That is, in this embodiment, interference caused by thermal expansion between the male and female rotors is absorbed by modifying the tooth profile of the female rotor.
ところで、雌ロータの修正歯形を求める場合、雌雄両ロ
ータの熱膨張を考慮するだけでは不十分である。すなわ
ち、この種の圧縮機においては、製作誤差等を考慮して
、運転時、ロータ歯面間に僅かの安全隙間が保持される
ように歯形設計される必要がある。この安全隙間は、一
般に歯直角断面において２０μｍ程度である。また、こ
の種のロータの歯面には、一般に焼付防止用コーティン
グが行なわれるが、上記雌ロータ歯形を修正するに当っ
てはこのコーティングの膜厚をも考慮しなければならな
い。このコーティング膜厚は一般に歯直角断面において
３５μｆｆｌ〜４５μＭである。By the way, when determining the corrected tooth profile of the female rotor, it is not sufficient to consider the thermal expansion of both the male and female rotors. That is, in this type of compressor, the tooth profile must be designed so that a small safety gap is maintained between the tooth surfaces of the rotor during operation, taking manufacturing errors and the like into consideration. This safety gap is generally about 20 μm in the normal cross section. Further, the tooth surfaces of this type of rotor are generally coated with an anti-seizure coating, but the thickness of this coating must also be taken into consideration when modifying the tooth profile of the female rotor. The thickness of this coating is generally 35 μffl to 45 μM in a cross section perpendicular to the teeth.
つまり、歯面間隙間は少なくともコーティングをし得る
だけの大きさとする必要がある。In other words, the gap between tooth surfaces needs to be at least large enough to allow coating.
表１に、各雌ロータ歯形点ａ２〜ｇ２における歯面間隙
間εの具体的な算出法を示している。以下、この表１に
ついて説明する。Table 1 shows a specific method for calculating the inter-tooth surface gap ε at each of the female rotor tooth profile points a2 to g2. Table 1 will be explained below.
く以下余白〉
＝１２−
表１において、Ｎｏ、 １ は、各雌ロータ歯形点
ａ２〜ｇ２について、軸直角断面における前記熱膨張干
渉量ε゛を示している。Ｎｏ、２は、各雌ロータ歯形点
ａ２〜ｇ２について、Ｎｏ、 １ に示した各熱膨
張干渉量ε゛を歯直角断面における熱膨張干渉量に換算
した値を示している。Ｎｏ、 ３は、歯直角断面にお
ける安全隙間、２０μｍｆｌをＮｏ。=12- In Table 1, No. 1 indicates the thermal expansion interference amount ε'' in the axis-perpendicular cross section for each of the female rotor tooth profile points a2 to g2. No. 2 indicates the value obtained by converting each thermal expansion interference amount ε'' shown in No. 1 into the thermal expansion interference amount in the tooth normal cross section for each female rotor tooth profile point a2 to g2. No. 3 is the safety gap in the tooth normal cross section, 20 μm fl.
２における各熱膨張干渉量に加算した値を示している。The value added to each thermal expansion interference amount in 2 is shown.
Ｎｏ、４では、歯直角断面におけるコーティング膜厚、
３５μｍ（雌雄両ロータの総合コーティング膜厚は３５
μｍ×２）を考慮し、Ｎｏ。For No. 4, the coating film thickness in the tooth normal cross section,
35μm (total coating thickness for both male and female rotors is 35μm)
μm×2), No.
３における各位が７０μｍ（３５μｗ×２）以下である
場合（Ｂロータ歯形点ｂ２．ｓ ｅ２ ｔ ’ｆ２
）につき、数値を７０μ石に修正している。Ｎｏ。If each part in 3 is 70 μm (35 μw×2) or less (B rotor tooth profile point b2.s e2 t 'f2
), the value has been revised to 70μ stone. No.
５は、Ｎｏ、４ における各位を換算した軸直角断面に
おける値、つまり求めるべき歯面間寸法ε（ε、〜ε７
）を示している。尚、歯面間隙間ε１〜ε７は！１１図
（ＩＩ）に表示している。5 is the value in the axis-perpendicular section obtained by converting each part in No. 4, that is, the tooth flank dimension ε (ε, ~ ε7
) is shown. In addition, the gaps between tooth surfaces ε1 to ε7 are! It is shown in Figure 11 (II).
さて、第１図（ＩＩ）における破線歯形、つまり、本実
施例において求めるべ柊雌ロータ修正歯形は実線歯形に
対して次のように表わされる。尚、修正歯形におけるａ
２゛〜ｇ２１は、実線歯形の各点ａ２〜Ｆｉ２に対応す
る点を示す。Now, the broken line tooth profile in FIG. 1 (II), that is, the modified tooth profile of the female rotor obtained in this embodiment is expressed as follows with respect to the solid line tooth profile. In addition, a in the modified tooth profile
2' to g21 indicate points corresponding to each point a2 to Fi2 of the solid line tooth profile.
・点ａ２゛〜ｂ２゛間は、点ａ２゛で歯面間隙間ε３を
、また点ｂ２“で歯面間隙間ε２を有する円弧である。- The area between points a2'' and b2'' is a circular arc having an inter-tooth surface gap ε3 at point a2'' and an inter-tooth surface gap ε2 at point b2''.
・点ｂ２゛〜ｃ２゛間は、点ｂ２゛で歯面間隙間ε２を
、また点０２′で歯面間隙間ε１ を有する円弧である
。- The area between points b2' and c2' is a circular arc having an inter-tooth surface gap ε2 at point b2' and an inter-tooth surface gap ε1 at point 02'.
・点ｃ２゛〜ｄ２１間は、点ｃ２゛で歯面間隙間ε５を
また点ｄ２で歯面間隙間ε１を有する修正創成曲線であ
る。- The area between points c2'' and d21 is a modified generation curve having an inter-tooth surface gap ε5 at point c2'' and an inter-tooth surface gap ε1 at point d2.
・点ｄ２１〜ｅ２１間は、点ｄ２１で歯面間隙間ε。- Between points d21 and e21, the gap between tooth surfaces is ε at point d21.
を、また点ｅ２゛で歯面間隙間ε７を有する円弧である
。is also a circular arc having an inter-tooth surface gap ε7 at point e2'.
・点ｅ２゛〜ｆ２１間は、点ｅ２で歯面間隙間ε１を、
また点ｆ２で歯面間隙間ε６を有する円弧である。・Between points e2 and f21, the gap between tooth surfaces ε1 is set at point e2,
Further, it is a circular arc having an inter-tooth surface gap ε6 at point f2.
・点１２〜ｇ２間は、点ｆ２で歯面間隙間ε６を、また
点ｇ２で歯面間隙間ε、を有する円弧である。- The area between points 12 and g2 is a circular arc having an inter-tooth surface gap ε6 at point f2 and an inter-tooth surface gap ε at point g2.
」二記歯形（破線歯形）の雌ロータおよび実線歯形の雄
ロータの各歯面に膜厚３５μｍの焼付防止用コーティン
グを行った−にで、これらのロータを圧縮機に組み込む
場合、各ロータのコーティング被膜の表面間隙間は、理
論上、点ｂ１〜１）２゛間、点ｅｌ〜ｅ２’間および点
ｒ、 −ｆ２’間において「０」になり、その他の点に
おけるコーティング被膜表面間隙間はｒＯＪ以上の適当
な寸法となる。従って、ロータ組付時においても、ロー
タのコーティング被膜間の干渉は生ぜず、ロータ組付作
業に何ら支障がない。そして、このようにしてロータの
組付を終えた圧縮機を運転し、ロータ温度が初期設定温
度２００℃になると、シールラインにおける各歯面間隙
間は、すべて理論的には安全隙間に等しくなり、以って
圧縮効率は非常に良好となる。尤も、各ロータ歯面には
コーティング被膜が存在するので、コーティング被膜間
隙間は」二記安全隙間よりも狭くなっている。The female rotor with tooth profile 2 (dashed line tooth profile) and the male rotor with solid line tooth profile are coated with anti-seizure coating with a film thickness of 35 μm on each tooth surface.When these rotors are assembled into a compressor, Theoretically, the gap between the surfaces of the coating film is 0 between points b1 and 1)2, between points el and e2', and between points r and -f2', and the gap between the surfaces of the coating film at other points is 0. is an appropriate dimension equal to or larger than rOJ. Therefore, even when the rotor is assembled, there is no interference between the coatings of the rotor, and there is no problem in the rotor assembly operation. Then, when the compressor with the rotor assembled in this way is operated and the rotor temperature reaches the initial setting temperature of 200°C, the gaps between each tooth flank in the seal line are theoretically equal to the safety gap. , thus the compression efficiency is very good. However, since there is a coating on each rotor tooth surface, the gap between the coatings is narrower than the safety gap described in ``2''.
第１，２図は本発明の実施例を示し、第１図（））は雌
雄ロータの南面間隙■肝０．１の想定常温時実線歯形と
雌ロータの求められるべき破線歯形を示す要部軸直角断
面説明図、第１図（ＩＩ）は第１図（■）における実線
雌ロータ歯形と破線雌ロータ歯形間の歯面間隙間を示す
要部断面説明図、第２図（Ｉ）〜（Ｖｌ）は第１図にお
ける実線歯形の両ロータを回転された場合の各回転角に
おける歯形干渉を示すための説明図、第３図（１）、
（ＩＩ）は雄ロータおよび雌ロータの熱膨張量を一般的
に示す説明図である。Figures 1 and 2 show an embodiment of the present invention, and Figure 1 () is the main part showing the assumed normal temperature solid line tooth profile of the male and female rotors with a south face gap of 0.1 and the dashed line tooth profile that should be determined for the female rotor. Fig. 1 (II) is an explanatory cross-sectional view perpendicular to the axis, and Fig. 1 (II) is an explanatory cross-sectional view of the main part showing the gap between the tooth flanks between the solid line female rotor tooth profile and the broken line female rotor tooth profile in Fig. 1 (■), and Fig. 2 (I) - (Vl) is an explanatory diagram showing the tooth profile interference at each rotation angle when both rotors with the solid line tooth profile in FIG. 1 are rotated, FIG. 3 (1),
(II) is an explanatory diagram generally showing the amount of thermal expansion of the male rotor and the female rotor.
Claims (1)
塗布してなるスクリュー圧縮機のスクリューロータにお
いて、 常温において歯面間隙間が［０」となるように配置した
雌雄ロータを想定するとともに、運転時に生ずる熱膨張
に起因する上記両ロータの軸直角断面歯形における各歯
形点の干渉量を算出し、該各歯形点の干渉量に、運転時
に最低限与えるべき安全隙間を加算した値を常温時の両
ロータ間における各歯形点間の歯面間隙間として保持す
るように両ロータ歯形を決定してなることを特徴とし、
かつ上記歯面間隙間が上記両ロータの焼付防止コーティ
ング被膜の総軸断面膜厚より小さい歯形点については、
上記歯面間隙間を上記焼付防止コーティング被膜の総軸
断面膜厚と等しい値にすることを特徴とするスクリュー
圧縮機のスクリューロータ。[Claims] 1. In a screw rotor for a screw compressor in which an anti-seizure coating film is applied to the surface of each male and female rotor, the male and female rotors are arranged so that the gap between the tooth surfaces is [0] at room temperature. Assuming this, the amount of interference of each tooth profile point in the axis-perpendicular tooth profile of both rotors due to the thermal expansion that occurs during operation is calculated, and the minimum safety gap that should be provided during operation is added to the amount of interference at each tooth profile point. The tooth profile of both rotors is determined so that the value obtained is maintained as the gap between tooth surfaces between each tooth profile point between both rotors at room temperature,
And for tooth profile points where the gap between the tooth surfaces is smaller than the total axial cross-sectional film thickness of the anti-seize coating of both rotors,
A screw rotor for a screw compressor, characterized in that the gap between tooth surfaces is set to a value equal to the total axial cross-sectional film thickness of the anti-seizure coating.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP3082185A JPS61190184A (en) | 1985-02-18 | 1985-02-18 | Screw rotor of screw compressor |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| JP3082185A JPS61190184A (en) | 1985-02-18 | 1985-02-18 | Screw rotor of screw compressor |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| JPS61190184A true JPS61190184A (en) | 1986-08-23 |
| JPS6254998B2 JPS6254998B2 (en) | 1987-11-17 |
Family
ID=12314366
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| JP3082185A Granted JPS61190184A (en) | 1985-02-18 | 1985-02-18 | Screw rotor of screw compressor |
Country Status (1)
| Country | Link |
|---|---|
| JP (1) | JPS61190184A (en) |
Cited By (4)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH03189392A (en) * | 1989-12-18 | 1991-08-19 | Hitachi Ltd | Oilless type screw machine |
| US5401149A (en) * | 1992-09-11 | 1995-03-28 | Hitachi, Ltd. | Package-type screw compressor having coated rotors |
| US6986652B2 (en) | 1999-11-17 | 2006-01-17 | Carrier Corporation | Screw machine |
| WO2018054858A1 (en) * | 2016-09-21 | 2018-03-29 | Knorr-Bremse Systeme für Nutzfahrzeuge GmbH | Assembly of screws for a screw compressor for a utility vehicle |
-
1985
- 1985-02-18 JP JP3082185A patent/JPS61190184A/en active Granted
Cited By (8)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| JPH03189392A (en) * | 1989-12-18 | 1991-08-19 | Hitachi Ltd | Oilless type screw machine |
| US5401149A (en) * | 1992-09-11 | 1995-03-28 | Hitachi, Ltd. | Package-type screw compressor having coated rotors |
| US5613843A (en) * | 1992-09-11 | 1997-03-25 | Hitachi, Ltd. | Package-type screw compressor |
| US6986652B2 (en) | 1999-11-17 | 2006-01-17 | Carrier Corporation | Screw machine |
| US6988877B2 (en) | 1999-11-17 | 2006-01-24 | Carrier Corporation | Screw machine |
| US7153111B2 (en) | 1999-11-17 | 2006-12-26 | Carrier Corporation | Screw machine |
| WO2018054858A1 (en) * | 2016-09-21 | 2018-03-29 | Knorr-Bremse Systeme für Nutzfahrzeuge GmbH | Assembly of screws for a screw compressor for a utility vehicle |
| CN109790751A (en) * | 2016-09-21 | 2019-05-21 | 克诺尔商用车制动***有限公司 | The screw arbor assembly of screw compressor for commercial vehicle |
Also Published As
| Publication number | Publication date |
|---|---|
| JPS6254998B2 (en) | 1987-11-17 |
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