Several key problems in external gear pump

Created on 05.17

Several key problems in external gear pump

a. The overlap coefficient (degree) e of gear meshing must be greater than 1, that is, at least two pairs of gear teeth must mesh at the same time. Therefore, a part of the oil is trapped between the closed cavity formed by two pairs of gear teeth, which is also called trapped oil area. The trapped oil area is not connected with the high and low pressure oil cavities of the pump, and changes with the rotation of the gear, as shown in Figure C. From figure C (a) to figure C (b), the volume of trapped oil area V decreases gradually; from figure C (b) to figure C (c), the volume of trapped oil area V increases gradually. The decrease of trapped oil volume will cause trapped oil to be squeezed and overflow through the gap, which will not only produce high pressure, make the pump drive shaft and shaft bear additional periodic load, but also cause oil heating; when the trapped oil volume changes from small to large, local vacuum and cavitation will be formed due to no oil supplement, causing cavitation and strong vibration and noise. Figure B shows the change curve of trapped oil volume. The problem of trapped oil not only affects the working quality of gear pump,

It can also shorten its service life.

The common measure to solve the problem of trapped oil is to set unloading grooves (grooves) corresponding to the trapped oil area on the inner surface of the front and rear covers of the pump. In addition to the double rectangular structure symmetrically arranged relative to the gear centerline (Fig. C), there are also the double circular unloading groove symmetrically arranged relative to the gear centerline [Fig. D (a)] and double oblique cutting unloading groove [Fig. C (b)] and the thin strip unloading groove symmetrically arranged relative to the gear centerline [Fig. D (c)]. The characteristics are different, but the unloading principle is the same, that is, on the premise of ensuring that the high and low pressure cavities are not connected with each other, the trapped oil area is connected with the high pressure cavity (oil pressure port) when the volume is reduced, and with the low pressure cavity (oil suction port) when the volume is increased. For example, the double dotted line in Figure C shows a symmetrical double rectangular unloading groove. When the volume of the trapped oil area decreases, it is connected with the oil pressure chamber through the unloading groove on the left [figure C (a)], and when the volume increases, it is connected with the oil suction chamber through the unloading groove on the right [figure C (c)].

In order to ensure better unloading effect and avoid oil suction and pressure area collusion, the size of unloading groove (such as the width and depth of rectangular unloading groove or the diameter and depth of circular unloading groove) and the spacing between two unloading grooves should be appropriate. In general, the two unloading grooves of gear pump are often offset to the oil suction area and opened asymmetrically. As shown in Figure e, the spacing a (minimum closed dead volume) between the two grooves must ensure that the oil suction cavity and the oil pressure cavity can not collude with each other at any time. For the standard involute gear with modulus m (the pressure angle of dividing circle is a), a = 2.78m. When the unloading groove is asymmetric, B = 0.8m must be ensured on the side of the oil pressure cavity The slot width Cmin > 2.5m and the slot depth h ≥ 0.8m.

b. The main obstacle of high pressure gear pump is that there are many leakage ways, and it is not easy to solve by sealing measures. There are three main leakage ways in the external gear pump: the axial clearance between the two sides of the gear and the end cover; the radial clearance between the inner hole of the shell and the outer circle of the gear; the tooth surface meshing clearance of the two gears. The axial clearance has the greatest influence on the leakage, because the leakage area is large and the leakage path is short. The leakage can account for 75% ~ 80% of the total leakage. The larger the axial clearance is, the larger the leakage is, which will make the volumetric efficiency too low; if the clearance is too small, the mechanical friction loss between the gear end face and the pump end cover will increase, which will reduce the mechanical efficiency of the pump.

The solution to the leakage problem is to select appropriate clearance for control: generally, the axial clearance is controlled at 0.03 ~ 0.04mm; the radial clearance is controlled at 0.13 ~ 0.16mm. In medium high pressure and high pressure gear pumps, the automatic compensation method of axial clearance is generally used to reduce leakage and improve the volumetric efficiency of the pump. The automatic compensation of axial clearance is generally to add floating shaft sleeve (floating side plate) or elastic side plate between the front and rear end covers of the pump to compress the gear end face under the action of hydraulic pressure, so as to reduce the leakage through the end face in the pump and achieve the purpose of increasing the pressure. The floating shaft sleeve can be replaced at any time after wear.

The principle of automatic compensation of axial clearance is shown in Figure F. The two meshing gears are supported by sliding bearings or rolling bearings in the front and rear axle sleeves 4 and 2, which can float axially in the housing 1. The pressure oil is led from the pressure oil chamber to the outer end of the shaft sleeve and acts on the area A1 with a certain shape and size. The resultant force of the hydraulic pressure is F1 = a1pg, which presses the shaft sleeve to the end face of the gear, and its size is proportional to the output working pressure PG of the pump.

The hydraulic pressure on the end face of the gear acts on the inner end face of the shaft sleeve, forming a reverse thrust on the equivalent area A2, which is also proportional to the working pressure, that is, FF = a2pm (PM is the average pressure acting on A2).

When the pump is started, the floating shaft sleeve is close to the gear end face under the action of elastic element (rubber sealing ring or spring) elastic ft to ensure the sealing.

In order to ensure that the shaft sleeve can automatically stick to the end face of the gear under various working pressures and automatically compensate after wear, the pressing force FY (= ft) should be adjusted +F1) is greater than the reverse thrust FF, but FY is not allowed to be too much greater than FF. The ratio of pressing force to reverse thrust FY / FF depends on the [PV] value of shaft sleeve and gear material and mechanical efficiency, that is, in order to reduce friction loss, the value of remaining pressing force (FY FF) should not be too large, so as to ensure that proper oil film can be formed between shaft sleeve and gear, which helps to improve volumetric efficiency and mechanical efficiency. General

Fy/Ff=1.0～1.2                         (2-1)

In addition, it is necessary to ensure that the action lines of the pressing force and the reverse thrust coincide, otherwise the couple will be produced, which will cause the shaft sleeve to tilt and increase the leakage.

c. The radial force problem and its countermeasures when the gear pump is working, the radial force F acting on the bearing of the gear pump is composed of the radial force FP generated by the liquid pressure along the circumference of the gear and the radial force ft generated by the gear meshing, as shown in Figure G.

When the gear pump works, in the radial clearance between the gear and the inner hole of the shell, the liquid pressure distribution from the oil suction chamber to the oil pressure chamber gradually increases step by step, and the approximate distribution curve of the liquid pressure is shown in Fig. G. The radial force FP produced by the liquid pressure on the driving gear and driven gear is exactly the same, and its direction is vertical and downward to the oil suction chamber. The radial force ft generated by gear meshing on driving gear and driven gear is approximately equal, but the direction is different. According to the radial force FP generated by the liquid pressure around the gear and the radial force ft generated by the gear meshing, the approximate calculation formula of the resultant force F1 of the radial force on the driving gear and the resultant force F2 of the radial force on the driven gear can be obtained

F1=0.75△pBDe                           (2-2)

F2=0.85△pBDe                           (2-3)

Where △ P -- pressure difference between inlet and outlet of gear pump;

B -- tooth width of gear;

De -- diameter of addendum circle of gear.

Obviously, the resultant force F2 of the driven gear is larger than that F1 of the driving gear. Therefore, when the specifications of the bearings on the driving wheel and the driven wheel are the same, the bearings on the driven wheel wear faster. In order to make the service life of the two bearings equal or close, the pressure oil port can be offset to the side with small radial force, so as to make F2 ~ F1.

Because the radial force is unbalanced force, and the higher the working pressure is, the greater the radial unbalanced force is. When it is serious, the gear shaft will be deformed, and the oil suction port side of the shell will be scratched by the gear teeth. At the same time, the wear of the bearing will be accelerated, and the service life of the pump will be reduced. There are two common ways to reduce the radial unbalance force.

Method 1: reasonable selection of gear modulus m and tooth width b (B / M = 6-10 for low pressure gear pump and B / M = 3-6 for medium and high pressure gear pump) can reduce radial force without reducing volumetric efficiency.

Method 2: change the pressure distribution along the circumference, such as reducing the size of the pressure oil port of the pump, so that the pressure oil only acts on one tooth to two teeth, or setting oil groove (balance groove) on the cover plate or around the shaft sleeve to reduce the radial force. As shown in Fig. h, the balance grooves 1 and 2 on the cover plate are connected with the low pressure chamber and the high pressure chamber respectively to generate a hydraulic radial force corresponding to the oil suction chamber and the oil pressure chamber to balance the radial force.

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