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渣漿泵揚程特性曲線繪制條件
目前所有葉片泵的特性曲線H=f(Q)和N=f(Q)，只能通過試驗方法來獲得，為此要進行泵的標準試驗，還沒有公認的計算方法，用來估算很寬流量范圍內(nèi)泵的水力損失。
根據(jù)本篇第三章第三節(jié)中所述的確定不同流量時理論揚程Hτ= f(Q)和壓水室內(nèi)水力損失hora=f(Q)的方法，通過計算方法可以繪制渣漿泵在相當寬的流量范圍內(nèi)揚程特性曲線H= f(Q)。為此應(yīng)該給出:
(1)葉輪轉(zhuǎn)速。
(2)葉輪主要參敷:直徑D2,出口寬度b2，葉片出口角風，葉片數(shù)z，葉片在葉輪出口處的厚度8.
(3)壓水室主要參數(shù)：壓水室斷面寬度B，高度h，計算斷面和隔舌斷面的形式，壓力短管喉部斷面面積Pr隔舌安放角4 (參閱圖3-3-1)。
二、揚程特性曲線繪制方法
采用下列方法繪制特性曲線H= f(Q):
1)確定不同流量時理論揚程和Ht=f（Q）的關(guān)系圖線。為此計算下列一些流量值時包括泵特性曲線工作段的參數(shù)：
葉輪出口圓周速度ua=Dxn/60和角速度。=rn/30;
葉輪出口軸面速度cs=Q(nD.b.);
葉輪出口液流排擠系數(shù)業(yè),不僅考慮葉片厚度，而且要考慮脫流區(qū)的存在[根據(jù)公
式(3-2-9)];
葉片出口角修正量AB為
根據(jù)公式(3-2-15)確定每個流量點的理論揚程。
嚴格地說，Ht不是泵流量的函數(shù)，而是流過葉輪液體流量的函數(shù)：Q=Q+q.為了簡化計算，可以不考慮泄漏量，因為確定這種條件下HT的誤差遠小于計算本身的誤差。
在必要時，利用下列方法考慮泄漏量q。確定泄漏量q，假定它是泵流量的函效，q=f（Q）=Q(1-1)。式中，為了泵的容積效率，根據(jù)本章第三節(jié)資料，是針對最佳狀態(tài)確定的。
因此，開始時先求最佳狀態(tài)的泄露量ga，然后根據(jù)下式計算它在其他狀態(tài)時的值
考慮泄漏量時，利用下列方法給制H-f(0)關(guān)系曲線：繪出曲線Ht=f（Q）的每一個點，沿著水平方向小流量側(cè)移動對應(yīng)給定工作狀態(tài)的 q值。
(2) 確定對應(yīng)最高水力效率的狀態(tài)。如果最佳參數(shù)Qam和Har給出，那么可以根據(jù)這些參數(shù)計算泵對應(yīng)最佳工作狀態(tài)的比轉(zhuǎn)速。
如果泵的最佳參數(shù)沒有給出，那么求出流量Q-.此值是計算壓水室螺旋(環(huán)形)泵體內(nèi)損失時所必需的。
為了確定Qmzx，從圖線Ht=f(Q)坐標原點繪制射線，根據(jù)下列關(guān)系式確定射線與流量直線的傾斜率
這條射線與直線Ht=f(Q)的交點就確定了最高水力效率狀態(tài)(這種方法的理論報據(jù)參同本箱第三章第四節(jié))。由這種方法求出的Qnm值與由一些渣漿泵平衡試驗數(shù)據(jù)所得到的這個流量值對比結(jié)果指出，試驗和計算值具有很好的一致性(誤差為5%以下)。同時，流量試驗值通常好于計算值，這可以用壓水室計算斷面上速度變化規(guī)律與Ro.=常數(shù)偏差不大來說明。
(3)對與第一節(jié)相同的流量，確定壓水室內(nèi)的損失，它是泵壓水室螺旋(環(huán)形)泵體內(nèi)和壓力短管內(nèi)的損失之和。
根據(jù)上面所確定的值，利用本篇第三章第三節(jié)(參閱計算例子)所述的方法確定壓水室(螺旋泵體)內(nèi)的損失。
(4)確定葉輪內(nèi)的水力損失，因為葉輪內(nèi)的水力損失在很寬流最范圍內(nèi)近似為常數(shù)，只有在很小流量狀態(tài)才開始增大，所以在繪制揚程特性曲線時，在整個討論流量范圍內(nèi)采用hx=常數(shù)。hx值是根據(jù)公式hx-Hr.amr(1一m加)針對最佳狀態(tài)(Qar) 確定的，根據(jù)葉片數(shù)z選取葉輪的水力效率(參閱本箱第四章第一節(jié))。
(5)計算泵過流部分總的水力損失: hu=hx + hurs.因為進口短管內(nèi)的水力損失不大，這里不予考慮(由于進口處的速度低和吸入短管形式——漸縮式或者圓簡式，從水力損失觀點看是有利的)。
(6)求泵的揚程H=H-ha.為了估算所提出方法的精度，就要計算20臺泵在很寬流量內(nèi)的揚程。吸入短管直徑在125-70mm之間變化，比轉(zhuǎn)速在70~200范圍內(nèi)變化，過流所面尺寸，壓水室形狀也在很大范圍內(nèi)變化，葉片數(shù)在2~4之間變化。
在接近或者大于最佳流量時，揚程計算值和試驗值近似一致(差值為5%~8%). 在Q≤(0.4~0.5)Qar時，發(fā)現(xiàn)揚程值離散度較大，在個別情況下可達到10%~12%。出現(xiàn)這種情況的原因是在確定壓水室內(nèi)的水力損失不夠精確，因為在很小流量時，液體從計算斷面和隔舌之間的區(qū)域已經(jīng)回流到葉輪內(nèi)，此外，在很小流量時，嚴格地說葉輪內(nèi)的水力損失不是常數(shù)，它已開始增大，目前還不能估算葉輪內(nèi)這種增大的水力損失，因此，對于比轉(zhuǎn)速n=70~200，流量Q=(0.51~1.5)Qamn的渣漿泵，可以推薦上述繪制揚程特性曲線的方法。
在確定實型泵揚程特性曲線時，采用模型泵特性曲線相似系數(shù)換算方法，一般考慮所謂尺寸效應(yīng)，其理由如下。當泵的尺寸增大時，由于過流部生分流道表面相對租糙減小，可
以觀察到水力損失有所降低，即提高了水利效率。這種情況將使實型泵揚程比相似換算得到的揚程有所提高。對不同尺寸(比轉(zhuǎn)速相同和特征系數(shù)kh2的流道斷面相對尺寸）的相似渣漿泵揚程特性曲線分析表明，實際上沒有觀察到由于尺寸效應(yīng)而使揚程提高。這是由于液體在渣漿泵流道——葉輪和壓水室內(nèi)流動時主要損失，不是水力摩擦損失，而是混合損失。大家知道，這種損失與粗糙度無關(guān)。因此，在用計算方法繪制渣漿泵揚程特性曲線時，可以不考慮尺寸效應(yīng)。在泵的尺寸增大時，機械效率提高，起碼泵的容積效率提高。渣漿泵廠家
Drawing Conditions of Head Characteristic Curve of Slurry Pump
At present, the characteristic curves H = f (Q) and N = f (Q) of all vane pumps can only be obtained by test method. For this reason, there is no accepted calculation method to estimate the hydraulic loss of pumps in a wide flow range.
According to the method of determining theoretical head H_= f(Q) and hydraulic loss hora=f(Q) in pressurized water chamber at different flow rates described in Section 3 of Chapter III of this paper, the head characteristic curve H= f(Q) of slurry pump in a fairly wide flow range can be drawn by calculation method. To this end, it should be given that:
(1) Speed of impeller.
(2) The impeller is mainly applied to: diameter D2, outlet width b2, blade outlet angular wind, blade number z, blade thickness 8 at the outlet of the impeller.
(3) Main parameters of water chamber: width B, height h of water chamber section, form of calculation section and tongue section, throat section area of pressure short pipe Pr tongue placement angle 4 (see Figure 3-3-1).
2. Drawing Method of Head Characteristic Curve
Characteristic curves H= f (Q) are plotted by the following methods:
1) Determine the relationship between theoretical head and Ht=f(Q) at different flow rates. For this purpose, the following flow values are calculated, including the parameters of the working section of the pump characteristic curve:
The circumferential velocity of impeller outlet UA = Dxn/60 and angular velocity. =rn/30;
The impeller outlet velocity cs=Q(nD.b.);
In the field of liquid flow exclusion coefficient at impeller outlet, not only the thickness of blade but also the existence of the stripping zone should be considered.
Formula (3-2-9)];
The blade outlet angle correction AB is
According to the formula (3-2-15), the theoretical head of each flow point is determined.
Strictly speaking, Ht is not a function of pump flow, but a function of liquid flow through impeller: Q = Q + Q. In order to simplify the calculation, leakage can be ignored, because the error of determining HT under this condition is far less than the error of calculation itself.
If necessary, the following methods are used to consider the leakage Q. Determine the leakage q, assuming that it is the function of pump flow, q = f (Q) = Q (1-1). In order to improve the volumetric efficiency of the pump, according to the information in the third section of this chapter, it is determined for the best state.
Therefore, the leakage GA in the best state is calculated at the beginning, and then its value in other states is calculated according to the following formula.
When considering the leakage, the H-f (0) curve is drawn by the following method: each point of the curve Ht = f (Q) is plotted, and the Q value corresponding to the given working state is moved along the small flow side in the horizontal direction.
(2) Determine the state corresponding to the highest hydraulic efficiency. If the optimum parameters Qam and Har are given, then the specific speed of the pump corresponding to the optimum working state can be calculated according to these parameters.
If the optimum parameters of the pump are not given, then the flow Q-. This value is necessary for calculating the internal loss of the screw (annular) pump in the pressurized water chamber.
In order to determine Qmzx, a ray is drawn from the origin of the graph line Ht=f(Q) coordinate, and the slope of the line between the ray and the flow is determined according to the following relations.
The intersection of the ray and the line Ht=f(Q) determines the state of maximum hydraulic efficiency (the theoretical report of this method is referred to in Section 4 of Chapter 3 of this box). The Qnm value calculated by this method is in good agreement with the flow value obtained from some slurry pump equilibrium test data (the error is less than 5%). At the same time, the flow test value is usually better than the calculated value, which can be explained by the small deviation between the velocity variation law calculated by the pressure chamber and the Ro. = constant.
(3) For the same flow rate as the first section, the loss in the pressure chamber is determined, which is the sum of the loss in the screw (annular) pump body and in the short pressure tube of the pump pressure chamber.
According to the values determined above, the loss in the pressurized water chamber (screw pump body) is determined by using the method described in Section 3 of Chapter III (refer to calculation examples).
(4) Determine the hydraulic loss in the impeller, because the hydraulic loss in the impeller is approximately constant in the widest range of flow, and increases only in the very small flow state. So when drawing the head characteristic curve, HX = constant is used in the whole range of flow discussed. The HX value is determined according to the formula hx-Hr.amr (1-m plus) for the optimal state (Qar), and the hydraulic efficiency of the impeller is selected according to the number of blades Z (see Section 1 of Chapter IV of this box).
(5) Calculate the total hydraulic loss of the flow passage part of the pump: Hu = HX + hurs. Because the hydraulic loss of the inlet short pipe is not too large, it is not considered here (because of the low speed at the inlet and the form of the inhalation short pipe - progressive or circular form, it is advantageous from the viewpoint of hydraulic loss).
(6) Calculate the pump head H = H-ha. In order to estimate the accuracy of the proposed method, it is necessary to calculate the head of 20 pumps in a wide flow rate. The diameter of the suction tube varies from 125 mm to 70 mm, the specific speed varies from 70 to 200, the size of the overflow surface and the shape of the water chamber also varies in a wide range, and the number of blades varies from 2 to 4.
When approaching or exceeding the optimal flow rate, the calculated value of head is approximately the same as the experimental value (the difference is 5%~8%). When Q < 0.4~0.5 Qar, it is found that the dispersion of head value is large, and in some cases it can reach 10%~12%. The reason for this is that the hydraulic loss in the pressurized water chamber is not accurate enough, because when the flow rate is very small, the liquid has flowed back into the impeller from the area between the calculated section and the tongue separator. In addition, when the flow rate is very small, the hydraulic loss in the impeller is not a constant, it has begun to increase, and it is not yet possible. This increased hydraulic loss in impeller is estimated. Therefore, for slurry pumps with specific speed n=70~200 and flow Q=(0.51~1.5)Qamn, the above method of drawing head characteristic curve can be recommended.
When determining the head characteristic curve of a real pump, the method of converting the similarity coefficient of the model pump characteristic curve is adopted. The so-called size effect is generally considered.
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