GB2036873A

GB2036873A - Silencing Fluid Pumps and Motors

Info

Publication number: GB2036873A
Application number: GB7846823A
Authority: GB
Original assignee: South Western Industrial Research Ltd
Current assignee: South Western Industrial Research Ltd
Priority date: 1978-12-01
Filing date: 1978-12-01
Publication date: 1980-07-02
Also published as: GB2036873B

Abstract

A pump, or motor, e.g. of the gear or swash-plate type, has an expansion chamber 22 for the outgoing working- fluid, the fluid impedance of the average volume subject to the outlet pressure or the distance from the centre line of the chamber to the farthest wall of the chamber being such that formation of standing waves in said fluid is obviated partially or wholly. <IMAGE>

Description

SPECIFICATION Improvements Relating to Noise Reduction in Pumps and Motors This invention relates to noise reduction in pumps and motors and especially, though not exclusively, in pumps and motors of the positive displacement and centrifugal types. When these devices are employed in hydraulic or pneumatic systems they produce pressure fluctuations in the fluid which cause fluid borne noise. These fluctuations are propagated at the speed of sound within the system and may be partially or wholly reflected from system components, such as bends, tees, valves, etc. Such multiple reflections result in a standing pressure wave within the system, the pressure at any position being a function of both the characteristics of the pump or motor and the system.This pressure variation can excite mechanical vibrations in the system components, lines and supporting structure which, in turn, generate air borne noise. The object of this invention is to enable pumps and motors to be designed in such a way as to reduce the pressure fluctuations generated in the system and thereby reduce the air borne noise radiated from the system.

It has now been discovered that the pressure variations and vibration in the system may be influenced considerably by the effective impedance of the pump or motor. The mathematical basis for this will be referred to below, but in general it can be stated as a fair approximation that the impedance is inversely proportional to the internal dead volume of the unit. In simple terms therefore, the impedance is reduced by increasing the dead volume. However, there are theoretical and practical limits to the dimensions of the dead volume and an important factor is the avoidance of standing waves. These tend to cause resonance and increase vibration which are, of course, most undesirable. The unit should, therefore, be designed to eliminate or reduce resonance at the frequencies of particular interest, particularly at the upper end of the working frequency range.

The invention is applicable both to pumps and motors since both can have a considerable effect on a fluid pressure system connected thereto. In the following specification, however, the invention is described for convenience with reference to a fluid pump, and except where the context requires otherwise this should be understood to include the use of a motor.

The invention is particularly applicable to pumps (and motors) having at least one pump element movable within a working volume such that fluid pulsations are produced at a frequency which depends upon the number of elements and the speed at which the unit is driving. For convenience this is referred to as the fundamental pumping frequency. Normally each unit is designed to operate over a selected speed range and the frequency of major importance is the maximum continuous rated pumping frequency.

Broadly stated, in one aspect the invention consists in a fluid pump having at least one pumping element movable in a working volume and including an expansion chamber or volume included with the working volume and with the outlet from the pump, the arrangement and dimensions being such that the fluid impedance of the dead volume including the working volume and the expansion volume will effectively reduce or eliminate standing waves within the working volume or dead volume at the maximum designed fundamental pumping frequency.

Broadly stated, from another aspect, the invention consists in a fluid pump having at least one pumping element movable in a working volume, a pump outlet communicating with the working volume, an expansion chamber communicating with the working volume and the outlet, the distance from the centre line of the chamber inlet or outlet port to the furthermost surface of the chamber being such that standing waves in the chamber are eliminated or reduced at the maximum designed fundamental pumping frequency of the pump.Preferably the expansion chamber is dimensioned such that its length (Ic) in the direction of the central axis of the chamber outlet does not exceed 3a nr preferably 2a l nr and its transverse dimension (y,) measured from the central axis of the chamber inlet or outlet port to the furthermost surface of the chamber, does not exceed 3a l nr preferably 2a nr where a=the speed of sound in the fluid (m/s), n=rotational speed of the pump or motor (rev/min) and r=number of pumping elements. Preferably, the minimum transverse dimension (y,) is about 0.75xthe diameter (dc) of the chamber outlet port.

The ratio of chamber length (Ic) to the diameter (dc) of the chamber outlet for a minimum condition should preferably not be less than 0.2.

In a preferred embodiment, the expansion chamber has a circular transverse cross-section with the axis of the chamber outlet aligned with the centre of the circle, but other shapes can be used according to the design of the pump or motor. For example, semicircular, elliptical or arcuate cross-section chambers may be used.

Also, the chamber inlet and outlet connections may be either "in-line" or "off-set".

The expansion chamber may, if desired, be attached to the pump or motor flange i.e. it can be an external chamber. In this case, the distance from the pump or motor flange to the inlet of the chamber is preferably less than 0.25 m.

The invention may be performed in various ways and some preferred embodiments, with several modifications, will now be described with reference to the accompanying drawings, in which Figure 1 is a diagrammatic representation of a positive displacement pump, hydraulic pipe and valve system.

Figure 2 is a graph showing the variation of the logarithm of the modulus of the internal impedance of a pump, log(lZsl) with the logarithm of the frequency, (log w) for a positive displacement pump.

Figure 3 shows cross-sections through different shapes of expansion chambers.

Figure 4 is a vertical section through an external gear pump having an expansion chamber.

Figure 5 is a vertical section through an axial piston swash plate pump having an expansion chamber.

The mathematical theory behind the invention can be explained by referring to the arrangement of Figure 1, which shows a positive displacement pump 10, hydraulic pipe 11 and valve system 12.

The pump is driven at a constant speed by a prime mover and delivers fluid under pressure into a rigid pipe line terminated by a valve. The pump may be considered as a flow generator producing a flow ripple Q5 which is modified by the internal impedance Zs of the pump. Although the flow ripple is generally a complex periodic waveform it can be shown by means of Fourier Analysis to consist of a steady level together with dynamic components at integer harmonics of pumping frequency.

By analysing each harmonic frequency in turn, and employing wave propagation theory, it may be shown that the pressure Px at any point x along the system is related to the pressure at the pump flange P0 by the expression:

where y is the wave propagation constant for the pipe line.

PT is the reflection coefficient at the termination valve, and L is the length of line.

In order to reduce noise levels, it is necessary to reduce Px.

The vector quantity,

is dependent purely on the system parameters and is independent of the source. However Pxt and hence the noise levels, can be reduced for a particular system by reducing the pressure P0 at the pump flange.

It may be shown that this pressure, PO, is dependent upon both system and pump parameters, such that:

where ZE is the entry impedance of the system at the frequency of interest, and is generally an unknown quantity. In most cases the pressure P0 can be reduced by reducing the internal impedance of the pump is As PO is dependent upon both the pump impedance (Zs) and the system impedance (ZE) then the relative magnitudes of the vector quantities Zs and ZE must be considered.

Rewriting equation (2) as:

(1) If ZE Zs then Pc=QsZs, and the fluid borne noise levels can be reduced by reducing Zs.

(2) If ZE=ZS then Po=ZsZs/2, and again the fluid borne noise levels can be reduced by reducing Zs (3) As ZEOZ5 then P,-too, and any change in Zs will reduce the fluid borne noise levels.

(4) If IZEX Z51 then Pc=OsZE, and Zs has no effect.

(5) As ZERO, PO < O, and there is no room for any improvement.

For a particular hydraulic system the entry impedance ZE is likely to be unknown. However, for the wide range of entry impedances considered above, a reduction in the fluid borne pressure ripple can be achieved in many cases by reducing the internal impedance of the pump. The same argument can also be applied to a motor.

It has been found that in the case of plane wave propagation the internal impedence of a positive displacement pump can be determined from the equation:

where a is the speed of sound in the fluid.

p is the density of the fluid.

w is the frequency of interest.

v is the internal or dead volume of the unit, i.e.

the average volume which is exposed to the output pressure.

K, R and I are coefficients dependent on the internal geometry of the unit.

In some cases, it has been found sufficient to use a simplified form of equation (3) which gives,

The characteristic predicted by this equation as a function of frequency, is shown in Figure 2.

At low frequencies, in the range 0 to w1, where w1 is about 15 , 1 tan a is approximately equal to wi a and equation (4) can be simplified further to give:

At frequencies above , wave propagation effects occur within the body of the unit and the characteristic shows the multiple resonances and anti-resonances at frequencies predicted by equation (4).

Anti-resonance occurs at a frequency of 7ta 21 and repeats at frequencies given by a 7r (+n7r), 12 for n=1, 2... Resonance occurs at a frequency of 7ta and repeats at frequencies given by a I The expansion chamber used in the invention serves to reduce the pump impedance Zs, thereby reducing the noise level.

Inspection of equations (4) and (5) shows that the magnitude of the internal impedance, Zs, may be reduced by increasing the internal volume v of the pump. The method by which this is achieved, however, is critical. If the volume is increased by positioning an expansion chamber between the flow generating mechanism and the outlet port of the pump, the dimensions of the chamber must be such that they do not invalidate plane wave propagation theory or lower the resonant frequencies shown on Figure 2. The former limits the largest cross-sectional dimension of the chamber and the latter limits the length of the chamber.

The length of the chamber must be sufficiently short to prevent the build up of standing waves, which increase the noise level, and the chamber should also have a restricted outlet to create reflections.

In order to meet these requirements, the invention utilises an expansion chamber attached to or situated within the body of the pump or motor. For bi-rotational units the chamber can be fitted at both inlet and outlet ports. The chamber may be of circular cross-section with concentric connections (provision being made for an air bleed). However, other shapes can be used which are compatible with a particular pump or motor design. Several possible chamber cross-section shapes are illustrated in Figure 3. Figures 3(i) and 3(j) illustrate "in-line" and "off-set" chamber connections respectively. The small circle in each of the other shapes indicates the position of the chamber outlet.

Figure 4 shows an external gear pump 20 having a pair of rotary gears 21 and an expansion chamber 22 located between the gears 21 and the outlet 23. The chamber 22 has a crosssectional shape as shown in Figure 3(a), but other shapes including the Figure 3(b) arrangement, with an eccentric chamber outlet, may also be used. The maximum allowable dimensions of such an expansion chamber can be determined from a knowledge of the rotational speed of the prime mover, the number of pumping elements, i.e. gear teeth, pistons or vanes, and the speed of sound in the fluid.

Hence, it can be shown that the following dimensions should not be exceeded, 3a lc < (6) nr where ic is the length of the chamber (m) a is the speed of sound in the fluid (m/s) n is the rotational speed of unit (rev/min) r is the number of pumping elements and 3a Yc < (7) nr where y, is the transverse dimension from the centre line of the chamber inlet or outlet port to the furthermost surface of the chamber.

Preferably 2a c < nr and 2a Yc < nr The minimum value of y, is about 0.75xdiameter (dc) of the outlet port, and ic should not be less than c 0.2 dc Figure 5 shows an axial piston swash plate pump 30 having an expansion chamber 31 between the moving parts 32 of the pump 30 and the restricted outlet 33. The cross-section of the chamber 31 in this case may be as shown in Figure 3(c), (d), (e) or (f).

As most pumps and motors operate efficiently over a wide speed range, the above calculations are based upon the maximum continuous speed rating of the unit. Similarly, the speed of sound is calculated for maximum continuous pressure and maximum operating temperature conditions. By way of example, consider a 10 piston pump having a maximum continuous speed of 3000 rev/min. At maximum operating conditions a=1400 m/s.

Then 2x1400 c < 3000x10 i.e. lc < 93 mm, and 2x1400 Yc < 3000x10 i.e. y < 93 mm.

If the chamber were circular in cross-section, then for the concentric arrangement shown in Figure 3(a), the diameter should be less than 186 mm.

For the eccentric arrangement shown in Figure 3(b), the diameter should be less than 93 mm.

However, it is likely that the size of the chamber will be limited by practical considerations such as allowable space and casing stress levels.

The same limitations imposed by equations (6) and (7) also apply to the sizing of the expansion chamber when positioned at the flange of a pump or motor.

The limits imposed on the dimensions of the chamber by equations (6) and (7) ensure that the expansion chamber does not produce a resonant condition at the critical noise frequencies. This leads to a substantial drop in noise levels.

Claims

1. A fluid pump having at least one pumping element movable in a working volume and including an expansion chamber or volume included with the working volume and with the outlet from the pump, the arrangement and dimensions being such that the fluid impedance of the dead volume including the working volume and the expansion volume will effectively reduce or eliminate standing waves within the working volume or dead volume at the maximum designed fundamental pumping frequency.

2. A fluid pump having at least one pumping element movable in a working volume, a pump outlet communicating with the working volume, an expansion chamber communicating with the working volume and the outlet, the distance from the centre line of the chamber inlet or outlet port to the furthermost surface of the chamber being such that standing waves in the chamber are eliminated or reduced at the maximum designed fundamental pumping frequency of the pump.

3. A fluid pump according to Claim 1 or 2 in which the expansion chamber is dimensioned such that its length (I) in the direction of the central axis of the chamber outlet does not exceed 3a nr and its transverse dimension (y,) measured from the central axis of the chamber outlet to the furthermost surface of the chamber does not exceed 3a nr where a=the speed of sound in the fluid (m/s), n=rotational speed of pump or motor (rev/min) and r=number of pump elements.

4. A fluid pump according to Claim 3 in which IC and y, do not exceed 2a nr

5. A fluid pump according to Claim 3 or 4 in which the minimum transverse dimension y, is about 0.75xthe diameter dc of the chamber outlet.

6. A fluid pump according to any one of Claims 3 to 5 in which the ratio of chamber length ic to the diameter dc of the chamber outlet for a minimum condition is not less than 0.2.

7. A fluid pump according to any of the preceding claims in which the expansion chamber has a circular transverse cross-section with the axis of the chamber outlet aligned with the centre of the circle.

8. A fluid pump including an expansion chamber or volume substantially in any of the forms described with reference to the accompanying drawings.