GB2311564A

GB2311564A - Gear machine, pump or motor

Info

Publication number: GB2311564A
Application number: GB9704365A
Authority: GB
Inventors: Dietmar Schwuchow; Karl-Heinz Hirschmann; Gisbert Lechner; Klaus Riedl
Original assignee: Robert Bosch GmbH
Current assignee: Robert Bosch GmbH
Priority date: 1996-03-29
Filing date: 1997-03-03
Publication date: 1997-10-01
Anticipated expiration: 2017-03-03
Also published as: IT1290094B1; ITMI970595A1; DE19612497B4; FR2746863B1; GB9704365D0; FR2746863A1; DE19612497A1; GB2311564B

Description

2311564 Gear machine

Prior art

The invention concerns a gear machine, pump or motor, with two gear wheels controlled in a rotatable manner in a housing, whose gear teeth are meshed together and which separate a pressure chamber and a suction chamber or discharge chamber, respectively, from each other, an instantaneous volume flow dV/d(p, of the hydraulic medium being displaced Mi relation to the angle of rotation o, of a driving gear wheel, and the meshed gear wheels having a gear ratio (P1/(P2.

Hydrostatic drive systems are employed in many branches of engineering. Various types of displacement machines are used to transform the hydraulic energy. Gear pumps and especially external gear pumps have been most widely used 'm fixed-displacement pumps. The mami reason for this is primarily their simple construction. On the one hand it results 'm high efficiencies and high operational reliability, even under difficult operating conditions, and on the other hand permits economical mass production. In use, due to the high, achievable energy density, the external gear pump offers in addition the advantage of a small space and low weight requirement.

Gear machines are generally constructed with at least one pair of gear wheels, which consists of two gear wheels with external teeth (external gear pump, Fig. 1) or one gear wheel with extohal teeth and one with internal teeth (internal. gear pump). One gear wheel with external teeth is driven and transmits the rotary motion to the second gear wheel having either internal- or external teeth. A.distinction is made between leading edges and trailmig edges, depending on the 2 direction of rotation. The leading edges transmit the rotary motion between the driving and the driven gear wheel. As is known m a gear pump, the medium being conveyed is conveyed in the tooth spaces from the suction chamber 4 to the pressure chamber 5. The tooth profiles which are meshed and in contact prevent the medium from flowing back from the pressure chamber to the suction chamber. Since the position of the point of contact, i.e. the contact points of the two tooth profiles during meshing, referred to the fixed housing, continuously varies, variations occur in the volume flow, resulting in pressure fluctuations in the pressure chamber which vary at the meshing frequencies.

Limitations in the use of external gear pumps often result from the noise characteristics, which are unfavourable compared to the internal gear pump, and which become noticeable particularly in combination with the hydraulic system. Apart from. the operating noise produced by the toothing, the fluctuations in the volume flow caused by the periodical meshing excite pressure changes and noise in the entire connected hydraulic system. In order to maintain and especially to extend the range of applications of external gear pwnps, an effective reduction in their generated noise is necessary. A starting point for this is an appreciable reduction in the pulsating volume flow which occurs during operation. Appropriate investigations in previous years concerned themselves with the optimization of the parameters of involute gear teeth and produced solutions such as the introduction of backlash-free toothing or the use of two offset pairs of gear wheels in a duo-pump.

In order to reduce noise in gear pumps, DE 4022500 AI and EP 0539396 B I propose a volume flow that is as constant as possible. This is based on the mathematical equation for the instantaneous volume flow dV/d(pl of a gear pump or gear motor:

3 dv = b. 2 + 1 2)_(,+i) 21 _ (1+ 1)g2 [(r., -ra2 ay (X) dl 2 1 Here r., and r.2 denote the radii of the addendum circles of the driving and driven gear wheel, b the width of the tooth face of gear wheels 1 and 2, ' = (P1/T2 the ratio between the driving gear wheel 1 and the driven wheel 2, rwI is the operating pitch circle of the driving wheel 1 and &Y the distance from the instantaneous point of contact Y to the pitch point C. The distance g, depends on the angular position (p, of the driving wheel.

If for the instantaneous volume flow dV/dT, in the above equation, the operating pitch circle rwI is replaced by the operating centre-to-centre distance aw between the gear wheels, then the mistantaneous volume flow is dV = b. 2 + 1 2 1 2 1)g21 [(ral r.) a + (y) 2 i+i W iay In this equation the width of the tooth face b, the radii of the addendum. circles r., and r,2, and the operating centre-to-centre distancea are fixed geometrical quantities. The distance gy of the instantaneous point of contact Y from the pitch point C varies between two extreme values at the meshing frequency. The fluctuation in the volurne flow is therefore repeated periodically at the meshing frequency. According to DE 40225 00 A 1 and EP 053 93 96 B 1, to compensate the fluctuation in volume flow the toothing of both gear wheels 1, 2 should be designed so that the only remaining variable quantity, namely the ratio i, is specified during meshing mi relation to the distance 9.1 so that the resulting volume flow fluctuation becomes zero. This is achieved by making the equation (y) constant.

4 However, difficulties arise in the practical representation of the gear wheel form calculated in this way. Relatively thick teeth and relatively narrow tooth spaces were obtained, which mi practice impede problem-free insertion of the respective teeth of the mating gear (Fig. 2). The formation of the tip of the tooth in particular would impede the sealing between the tip and the pump housing.

Advantages of the invention

Claims

In contrast the gear machine according to the invention, having the

characterising features of Claim 1, has the advantage that tooth profiles can be demonstrated in practice with relatively low volume flow pulsation. A relatively large tooth space, i.e. a large trailing edge space and, nevertheless, minimum noise generation due to pulsation at the same time, is achieved with a steep edge shape.

Further advantages and advantageous developments of the gear machine according to the invention are disclosed in the sub-claims and the description. Above all, the noise performance of the hydrostatic system, consisting of pump, circuit and load, is improved.

Drawing To explain the terms, a generally known gear machine is shown schematically Mi Figure 1. Figure 2 shows schematically the shape of the gear teeth and tooth spaces according to the prior art; Figure 3 likewise shows the backlash-free construction according to the invention. In practice the trailing edges of the driven wheel would have a small amount of backlash. Figure 4 shows a comparison of the conveyed volume curve dV/d(pl versus the distance &Y between a gear machine of the prior art and the invention, and Figure 5 shows a comparison of the ratio i versus the distance g., Description of the exemplary embodiment It is important for the invention that a tooth profile that can readily be used in practice is obtained for the gear wheels. In this connection, the influence of the volume flow should be taken into account and made to deviate from a constant volume flow in a specific manner. A suitable deviation which still produces a relatively small pulsation is obtained if the volume flow is specified according to dv v - - g. (1 - f (gy (Z) dT 1 2n Here Vi2n represents the maximum instantaneous conveyed volume. The volume flow is defined in relation to the distance g,,y.

f > 0 must be set for g,,, > 0.

This gives a conveying characteristic which has a maximum value in the centre of engagement (Y = C) and a minimum value at a change of meshing point. A volume flow fluctuation of 4 to 5 percent referred to the average conveyed flow is advantageous.

Equating the equations (y) and (z) gives the conditional equation v g - ( 1 - f(g",)) = b - [(r 2 + r 2) - 1 a 2 + g2 2n 2 al C i+i W IXY from which the gear ratio flinction 1 (gy) can be generated.

During the design, to ensure kinematic compatibility, moreover, the average ratio 6 at meshing must correspond to the ratio of the number of teeth of gear wheels 1 and 2.

Other boundary conditions are that both gear wheels must have whole numbers of teeth and no changes in gear ratio occur on transition to the next meshing. Furthermore, the accelerations and decelerations of the driven gear wheel 2 must be kept within limits. In order to ensure a continuous transfer of torque, the profile overlap must be greater than 1. It is therefore imperative that the instantaneous ratio of both meshed teeth must be equal in the double area of contact.

The toothing in Figure 3 has been shown as being backlash-free, but in practice is designed with adequate backlash. The trailing edges of the gear wheels 1, 2 have therefore been designed so that they have the same instantaneous ratio as the driving, i.e. sealing tooth profiles so that there is no interference with the transmission of the rotary motion in the event of possible contact between the trailing edges. But because of the backlash they have no influence on the pressure pulsation. The tooth profiles of the driven and the driving wheel are therefore asymmetrical.

In Figure 4, the continuous line shows the conveyed volume dV/dT, for f (g Y) 0, i.e. a constant voluine flow according to the prior art. The broken line shows the curve of the conveyed volume according to the invention.

Figure 5 shows the gear ratio function 1 (gy) on which the toothing calculation is based. It shows the relationship between the instantaneous ratio i and the distance g.y, which the toothing of a pump must fulfil. Here the continuous line again shows the curve according to the prior art and the broken fine that of the invention. For example, an instantaneous conveyed volume which decreases parabolically produces larger ratio values for identical g,,, values.

7 Here the ratio i is defined positively for pairs of external gears, so that the curves illustrated in Figure 4 and 5 are valid for external gear pumps.

However, the ratio of the tooth space width ew to the tooth thickness sw is critical for each gear wheel. If these ratios are in the region of 0.6 to 1.7, then a usable tooth form is obtained with justifiable pulsation. If the ratio is limited to 0. 9 to 1.3, the manufacture of the tooth form is further simplified. The form of the tooth tip, above all, facilitates improved sealing. A ratio of 1 should be aimed for to achieve an ideal profile geomeuy.

The ratio of the tooth space width ew to the tooth thickness, could also be expressed as ew or SW ew + sw ew + sw For an optimum gear machine design, the tooth space width ew should be in the region of 60 % to 40 % of the circular pitch (sum of e, + sw) or the tooth thickness s, in the region of 40 % to 60 % of the circular pitch, corresponding to 100 % 8 CLAIMS 1. A gear machine, pump or motor, with two gear wheels (1, 2) controlled in a rotatable manner in a housing (3), whose gear teeth are meshed together and which separate a pressure chamber (5) and a suction chamber or discharge chamber (4), respectively, from each other, an instantaneous volume flow dV/d(p, of the hydraulic medium being displaced in relation to the angle of rotation (p, of a driving gear wheel (1), and the meshed gear wheels (1, 2) having a gear ratio ' = TI /T2, the instantaneous ratio i being determined by the design of the toothing over the entire angle of rotation (p, of the driving gear wheel (1), so that the driven gear wheel (2) is operated at a continuously varying angular velocity which is repeated periodically according to the circular pitch, characterised in that the volume flow corresponds to the equation v b. 2 + 1 2 1 a 2 + g2 g.0 f(g.,)) Url _. r.2 W ay 2n 2 1 1+1 where b is the tooth face width, r.,, r.2 the radii of the addendum circles, aw the operating centre-to-centre distance and g,,, the distance from the instantaneous point of contact Y to the mistantaneous, pitch point C, and that the ratios of the tooth space width e, to the tooth thickness sw of each gear wheel (1, 2) are in the region of 0. 6 to 1. 7.
2. Gear machine according to Claim 1, characterised in that the ratios of the tooth space width ew to the tooth thickness sw are in the region of 0. 9 to 1.3.
3. A gear machine substantially as herein described with reference to the accompanying drawmigs.