an experimental investigation into the design and performance of hydrostatically loaded floating...

ELSEVIER Wear 209 (1997) 184-192

WEAR

An experimental investigation into the design and performance of hydrostatically loaded floating wear plates in gear pumps

E. Ko~ a,• C.J. Hooke b Engineering Faculty, ~.ukurova University. 01330 Balcali, Adana, Turkey

b Birmingham University, School of Manufacturing and Mechanical Engineering, Birmingham, England

Received 18 September 1996; accepted 24 January 1997

Abstract

The design principles and the performance of hydrostatically loaded wear (end) plates in three high-pressure external gear pumps were examined experimentally while the pumps were miming under typical operating conditions. The film thicknesses (clearances) between the gears and end closures (plates) in three different designs of floating plate gear pump were measured. It was found that at high delivery pressures all the plates were heavily tilted relative to the gears. The plates appeared to be tilted so that they occupied the total available end clearances in the units. © 1997 Elsevier Science S.A.

Keywords: Hydrostatically loaded; Floating wear plates; Gear pumps; Experimental

1, Introduction

The leakage across the sides of the gears depends upon the clearances that exist between the side plates (wear plates) and gears (Fig. 1 ). The leakage through the ends of the gears is typically the largest proportion of the total internal leakage in this type of pump [ I-3].

A number of different end plates have been designed in an attempt to minimize the leakage across the ends of the gears and many of these are in production today. While controlling the side leakage, the plates must provide an adequate fluid film between the surfaces. In floating end or wear plate designs the end plates are clamped on to the ends of the gears by hydrostatic loads acting on the rear of the plate. In this design movable plates are located between the front and end casings and the gears. The pressure balancing is achieved by feeding high pressure to the back of the plates. The design procedure here is to load the plate lightly from the rear towards the gear end faces under all running conditions. The major problem in this type of plate is maintaining an adequate film thickness under all conditions and ensuring that the plate balance is not upset, which would result in an unstable operation, i.e. tilting, lifting off or seizure. The pressure variation around the circumference of the gear is important and an-

* Corresponding author. Correspondance address: Cukumva University, Department of Textile Engineering. 01330 Balcali. Adaoa. Turkey. Tel: 0 322 3386456. Fax: 0 3223386126.

0043-16481971517.00 © 1997 Elsevier Science S.A. All fights reserved PII S0043-1648 (97) 00021-5

stable operation may occur if the pump inlet cavitates. The hydraulic balancing of the plates is a main factor to be con- sidered in operating successfully without becoming unstable under any conditions.

This paper presents the results of an experimental investigation into the behaviour of floating plates in gear pumps. The study forms part of a wider investigation into the lubrication and sealing of the end closures in gear pumps of different designs [4-7].

2. Tested pumps and clearance measurement procedure

2.1. Testedpumpsandmodif icat ions

The clearances between the gear ends and the floating wear plates were measured in three different designs of pumps under a wide range of operating conditions.

2.1.1. Pump ! Fig. 1 shows a general view of a pump (pump i) with

floating type end plates. Two bronze end-plates are fitted in the housing. These are designed in such a way that they can float axially in the clearance that is provided in the centre case. The design clearance in this type of pump is maintained by loading the floating end plates from the rear. The procedure here is to apply a load at the rear of the plate just larger than

E. Ko¢. C . Z Hooke / Wear 209 (1997) 184-192 185

M o v i e wear prate

iii.ii., Fig . I. A general v i e w o f a p u m p wi th f loat ing wea r plates ( p u m p 1 ).

the hydrostatic forces generated between the gear and end plate. To do this, the delivery and the other pressures from the unit are connected to pockets at the back of the plates so as to clamp the plate onto the gear end faces.

In this design of end-plates there are four separate pockets at the rear of the plate (Fig. 2). Pocket Pl, at the centre of the plate, is pressurized by the leakage between the gear and plate itself. Two bronze floating bush seals are used on the drive shaft directly after the needle roller beatings to seal the high pressure in that pocket. The actual pressures in these pockets are determined by the balance of the flows into the pocket under the teeth in the high-pressure region out to the low-pressure region and also any additional leakage through the bush seals. Pocket P2, under the high-pressure teeth, is pressurized directly from the delivery. Pocket P3, under the boost section of the unit, is connected directly to the boost, and the third pocket around the perimeter, P4, is connected to the middle of the sealing arc. These pockets are separated by robber seals.

The two central pockets of the two wear plates are connected together to avoid any load unbalance, which means that the behaviour of one plate is affected by the operation of

the other. The actual dimension of the central cavity is determined by the dimensions of the bearing housing onto which it fits. On the top surface of the plates there are silencing recesses machined on both delivery and boost sides of the plates in an attempt to relieve the oil in the meshing zone of the gears. The spaces beyond both of the bush-seals are connected to the boost via a non-mum valve. The driven gear shaft is hollow so as to maintain the hydraulic balance of the gear.

In order to accomodate the transducer it was necessary to modify the pump. The drive shaft was altered as shown in Fig. 3 (a) and the transducer cable was taken out of the pump through a hollow quill shaft [Fig. 3(b) ]. The quill shaft was fitted inside the modified shaft and protruding through the pump end cover. Two o-rings, one on surface A and the other on the groove provided (B), were titled on this shaft. A lip seal was installed in the end cover to prevent oil leakage. The oil passages in the quill shaft ensured that the original hydrostatic balance of the pump was unaltered. The quill shaft was used to drive the screened slip ring through which the transducer signals were conveyed to the rest of the measurement equipment. One small hole on the end cover just before the lip seal was drilled in order to drain leakage passing through the bush seal. The central housing section was modified to measure the mean pressures around the gear perimeter. A small hole was drilled on the end cover in order to measure the pressure in the region below the wear plate.

2.1.2. Pump 2 The pump was modified in exactly the same way that was

done in pump 1 since it is virtually identical to that pump. Pump 2 is bigger in size than pump ! and has two dowels in the centre case to locate the three pump cases accurately. The wear plates used in this pump have just one silencing recess as shown in Fig. 4. The pressure measurements around the perimeter of the gear and below the wear plate away from the drive end were taken.

A ~ 4

A ~

SECTION B-B

I F ig . 2. W e a r plate for p u m p I.

SECllON A-A

l ( a ) ~-,~'///////,,~I V/////////-~H//"/"/. A. • -~

. il

(b) ~ , a B

.~'~ f t U:Tofit ~ ¢

Fig. 3. Drive gear and modifications (pump I ).

186 E. Kof, CJ. Hooke I Wear 209 (1997) 184-192

tow

silencing r l i c~$

Fig. 4. Wear plate for pump 2.

2.1.3. Pump 3 The tested pump body is made from two sections: front

and end casings which are joined together by four bolts. Two dowels are used to hold them in position. The gear shafts are both supported by four needle roller bearings. The driven gear is hollow, ensuring that two ends of the shaft are identical hydraulically. Two aluminium thin wear plates are used in both sides of the gears to maintain the required clearance in order to control the side leakages.

The wear plate used in the pomp is shown in Fig. 5. There is a recess machined on each wear plate at the locations where the gears fully mesh [ Fig. 5 (a) ]. These are silencing grooves relieving the trapped oil in the meshing section of the gear teeth. There are also four small holes drilled through on the plates. These holes were provided to connect the top and bottom of the plate in the high-pressure region. These arrangements ensure that there is a pressurized region at the bottom of the plate that was separated from the low-pressure zone [Fig. 5(b)] . There is a single pressurized area sealed by an wring and backing ring (area A) connected directly to the delivery of the pump. The remaining area (area C) is connected to the suction. In addition to the main connection, there is a subsidiary connection through the plate in the sealing arc (holes B), presumably to improve the balance of the plate itself.

The centre of the plates are drained through the slots machined on them to the inlet side of the unit [slot S in Fig. 5(b)] . The spaces beyond the roller bearings are all inter-connected through the passages machined in front and end casings in order to balance the two ends of the shafts and ensure that the area C at the rear of the plates is connected to the suction side of the unit. The lip seals are also fitted in the drive side of the front case to complete the sealing arrangement.

The pump was modified in order to accomodate the displacement transducer. The driving gear was modified and the hollow quill shaft was built to take the transducer cable out. A lip seal was used on this quill shaft to seal the oil coming out of the pump body.

2.2. Clearance measurement procedure

The behaviour of the floating wear plates was investigated experimentally by measuring the film thickness between the gear ends and floating end plates, while the pumps were

up r ~

(a) A J "~Ecr~ A-A

/ O-r~

Back up

Fig. 5. Wear plate for pump 3.

running under different operating conditions. This was achieved by using capacitance type displacement transducers installed in the end faces of the gear teeth oftbe driving gears. One transducer per pump was used and the transducers were positioned on tooth centre lines at 82% (pump I), 83.4% (pump 2 ) and 79.6% (pump 3 ) of the gear outer radius. This transducer continuously monitored the clearance between the end face of the gear and the plate as the gear rotated and provided a record of its variation around the pump. The results were obtained with a data logging system and the details of it have been reported elsewhere [8-10] and no details will be given here.

3. Test results and discussions

3.1. Results and interpretation

Some of the results obtained are shown in Figs. 6-13. Each plot shows the clearance variation for two shaft revolutions except at the lowest speeds where, because of limitations on computer storage, a lower number of revolutions were recorded. The vertical distance from the horizontal axis rep- resents the film thickness, with each mark on the vertical axis

E. Kof, C.J. Hooke / Wear 209 f1997) 184-192 187

" ! I ! , ! °"+-" I

(a)

A+ _=it It / ° t iL-,

i.~-1 Revolution (b)

,o I" I

(¢) (d)

Fig. 6. Measured film thickness for pump I at 7 MPa. 1000 rpm.

SO0 rpm 1000 fpm (a) (b)

corresponding to 5 p.m. The relatively large clearance zone " A " in each plot marks the passage of the transducer over the silencing recess on the plate.

It may be noted that, in order to avoid any possible inlet cavitation and consequent variations in plate balance, the measurements were taken with an inlet boost of 0.35 MPa (50 psi) and all delivery pressures quoted are relative to this boost pressure. All results were obtained with a hydraulic oil having a viscosity of 30 mPa s at the pump inlet temperature.

The tests were carried out over a long period of time and some results were obtained after the pump had been stripped and reassembled. The results presented here are a typical selection from the tests. In some tests, the pumps were loaded gradually from the boost to full delivery and then unloaded

1500 rpm (c)

Fig. 7. Measured film thickness for pump I at 7 MPa.

in an attempt to investigate any hystefisis. It should be noted again that the pumps were stopped for a short period between each squence of test and run up to speed under the low- pressure condition and then loaded up to the required pressure level.

3.1.1. Experimental results f o r pump I Fig. 6 shows typical clearance variations obtained from

pump I. The large clearance zones in each curve, marked A in Fig. 6(a), correspond to the passage of the transducer across the two decon:pression recesses of the plate. The mid position between these two recesses corresponds to the passage of the transducer across the pump centre line (shown as a chain dotted line in the figure). It may be noted that the

188 E. Kof, C.J. Hook¢ / Wear 209 (1997) 184-192

71W!II" ' ' i1 !11 ' "

i i i = " : i

i ' No Load 3.5 MPalSO0 psi) 7 MPa(IOQO psi)

(a) (b) (c)

.o I I"

10.S MPa(1500 psi) 14 MPa(2000 psi) (d) (e)

Fig. 8. Measured film thickness for pump I at 1500 rpm.

'® II . - 4 l f /

~'D--1 Revotutioi'l----4~ (a)

,® ~i ,_HI _ ni i i

(b)

(c) (d) Fig. 9. Measured film thickness for pump 2 at 7 MPa, I000 rpm.

suction zone comes after the second decompression groove and the delivery zone occurs before the first decompression groove.

The main feature of each clearance curve is a once per revolution variation of clearance, with the maximum lying 90 ° before the pump centre line. It would appear from this that the plate is tilted relative to the gear, with the minimum film thickness lying in the low-pressure region and the maximum clearance in the high-pressure region. This tilt takes place about an axis passing through the gear centres.

The clearance curves demonstrate the clearance variation at 7 MPa ( 1000 psi) delivery pressure and 1000 rpm running speed taken from an arbitrary position throughout the test, including a number of strips and rebuilds. It may be seen that the traces are closely similar. The minimum clearance varies from 3 p,m in (a) to 6 pm in (c) and the maximum clearance varies from 58 I.LIn in (a) to 66 ixm in (b). It would appear

that the results are reasonably repeatable, but there appears to be some hysterisis effect operating in the system.

3. I. 1. i. Speed effect The effect of speed on the wear plate behaviour was invcs-

tigated for a range of delivery pressures. Fig. 7 shows the clearance variation at 7 MPa ( 1000 psi) delivery pressure and 500, 1000 and 1500 rpm rotational speed. It is seen that the traces are roughly same. The speed has no effect on the behaviour of the plate. The general shape of the curves does not change with speed.

3.1.1.2. Pressure effect The effect of delivery pressure on the wear plate behaviour

was investigated by carrying oat the tests at different pressures and speeds. Fig. 8 demonstrates the film thickness variation at no load (0.35 MPa boost pressure), 3.5 (500),

E. Ko~. C.J. Hooke / Wear 209 (1997) 184-192 189

//i

3,5 MPa(SQOpsi)-IO00 rpm (a)

A I •

3,5 MPa( 500 psi)- 1500rum (U)

7 ~lpa (I(XX) psi )-I 5~) rpm (c)

Fig. lO. Measured film Ihickness for pump 2.

3S MPa(SOOpsi)- 500 rpm

1Q5 MPa (1500 psi) -1500 rpm (¢)

Fig. I I. Measured

10.5 HPa( 1SO0 psi)-tO00 ~'pm (b)

14 Mpa( 2000 pSi)-500 r pm (d)

lilm lhickness for pump 2.

10.5 (1500) and 14 MPa (2000psi) delivery pressures at 1500 rpm running speed, it may be seen that at high pressures a flattened zone occurs in the low-pressure region and the sinusoidal pattern of the clearance is distorted. The plate is heavily tilted under all pressure conditions and there is a distortion in the low-pressure region. Some distortion appears to be present even at 3.5 MPa (500 psi) and it is quite obvious at 7 MPa ( 1000 psi). It becomes more noticable as the pressure is increased.

The measured minimum clearance reduced slightly in the low-pressure region with pressure. However, because of the distortion no attempt can be made to extrapolate the minimum clearance. The maximum clearance appears to remain about 50-80 p~m. The plate wa~ running with rough tilt at no load

condition, but increased delivery pressure even 500 psi resulted in a considerably increased tilt and a reduced minimum film thickness. The previous running conditions affect the behaviour of the plate. This type of behaviour is expected because of the high side loads on the wear plate and its close fit into the centre case and over the bearing. In interpreting these results it should be noted that total axial end float in both ends of the gear is in the order of 100-200 g,m.

3.1.2. Experimentalresultsforpump 2 Fig. 9 shows typical clearance variations obtained from

pump 2. This plate has a single decompression groove, marked A in the figure, lying on the high-pressure side of the pump centre line. The figure exhibits an apparent increase in

190 E, go~, C.J. Hooke I Wear 209 (1997) 184-192

S00rpm (a)

I

40 A [ a f

C A C

~1 R e v o l u t i o n ~

1500rpm (c)

Fig. 12. Measured film thickness for pump 3 at 7 MPa.

1000 rpm

(b)

40 [ B A I A l e

NoLoad (a)

14 NPa (20001~) (d)

e C A I B c

&516Pal 500ps~) (b)

! iv' I

21 MPa(3000psi) (e)

Fig. 13. Measured film thickness for pump 3 at I000 rpm.

clearance, marked B, in the low-pressure region. This would appear to be due to cavitation of the fluid between the gear end face and the plate. Cavitation results in a reduction in the dielectric constant and hence an apparent increase in clearance.

The figure shows the measured clearance variation obtained at 7MPa (1000psi) delivery pressure and 1000 rpm rotational speed at different times, but without stripping and rebuilding the pump. It may be seen that the traces demonstrate apparently two different operating modes of the wear plate. The first two curves are virtually identical. The traces in (c) and (d) are nearly identical except a shift in (d). It would appear that the transducer while taking the trace in (d) was displaced and the zero level was in error. Unfortunately the transducer failed before it could be checked

i i . c

7 MP.~ (1000psi) (c)

3.1.2.1. Operatingconditionseffect The pump was run under different delivery pressures and

speeds. Some typical results are given here. The pump appears to operate in two modes. In the first mode it runs with the plate nearly flat and parallel, although a small amount of tilt of the two adjacent surfaces of wear plate and gear is observed. Fig. lO demonstrates this first mode of operation. The traces included the cavitation occurring in the suction side of the plate (represented in the curves by a relatively large spike). The effects of pressure and speed on the behaviour are not consistent. Under all pressure conditions there is

whether the shift from figure (b) to (c) was a permanent one or was due to the transducer effect. However, the results in the two modes are reasonably repeatable.

E. Kof, C.J. Hooke / Wear 209 (19971184-192 191

local distortion occurring in the high-pressure zone around the silencing groove. The plate seems to be iewardly deformed by pressure (the increases in clearances around the silencing groove).

In the second mode of operation the plate seems to be running heavily tilted with the tilt over 100 p.m in most operating conditions. Fig. I 1 shows selected results. The plate is tilted in such a way that the minimum clearance occurs in the low-pressure region just after the silencing recess (in the section where the gears come out of meshing zone). The very high pressure, 14MPa (2000psi), Fig. I I(d) , moves the location of minimum film thickness towards the sealing arc, but again the minimum clearance occurs in low-pressure region. The cavitation occurs in most cases in the low-pressure region in the diverging section of the oil film. The measured minimum film thickness varies from 10 to 20 p.m and the measured tilt varies between 45 and 110 p.m with pressure. The speed does not have any significant effect.

Once the plate starts operating in this second mode it settles down and runs heavily tilted, and it does not seem to return to the first running mo~_e. The clearance v~ries from run to run and the pressure has a significant effect on the wear plate behaviour.

3.1.3. Experimental results forpump 3 Fig. 12 presents the clearance variations obtained from

pump 3. The main large clearance zones in each curve, marked A in the figure, correspond to the passage of the transducer across the decompression groove which is situated on the high-pressure side of the pump centre-line and the two short high clearance spikes, marked C in the figures, represent the passage of the transducer over two small holes drilled through the wear plate. Cavitation again appears to be present in the low-pressure region of this pump producing the apparent high clearance region denoted by B in the figures. The results presented in the figures correspond to the running conditions of 7 MPa ( 1000 psi) delivery pressure and 500, 1000 and 1500 rpm rotation speed, respectively. I: ig seen that the cavitation in the low-pressure region occurs just under the speed of 1000 and over. Generally the speed has very little effect on the behaviour of the wear plate.

3.1.3.1. Pressure effect Fig. 13 shows the clearance variation at no load, 3.5 ( 500 ),

7 (1000), 14 (2000) and 21 MPa (3000 psi) delivery pressure and 1000 rpm running speed. It may be seen that in this pump the clearances are somewhat smaller than those found in pumps I and 2 and it also appears that the plate itself is not flat, with irregularities of the order of 5 p.m being recorded. This non-flatness complicates the general clearance pattern, hut it may be seen that under no load conditions the plate lies roughly parallel to the end face of the gear (the apparent high clearance region caused by the fluid cavitation must be ignored in interpreting the results). Then as the delivery pressure is increased the plate appears to tilt relative to the end face of the gear, producing a minimum clearance in the high-

pressure region and a maximum clearance in the low-pressure region. Again the plate seems to tilt about an axis passing through the gear centre lines, but the direction of tilt is oppo- site to that of pumps ! and 2. In addition the pressure appears to distort the wear plate in the high-pressure region around the silencing groove. This results in a slight increase in clearances in that zone. A t'urtber increase in pressure, up to 21 MPa, results in increased tilt and minimum film thickness [Fig. 13(e) I.

Some hysterisis was found in the magnitudes of the tilts recorded under low delivery pressure conditions, with the highest magnitudes of tilts being measured after the pomp had previously run with a high delivery pressure. The maximum clearance variation found under no load condition was, however, relatively low, being less than I 0 p.m. At the highest delivery pressure for which results were obtained [Fig. 13(d) I the clearance variation recorded was between 40 and 50 ~Lm.

3.2. General observationsanddiscussions

The running speed appeared to have little effect on the clearances measured, with virtually identical results being obtained at 500 and 1500 rpm for all delivery pressures in all the pumps tested. For instance, in pump 3 m 7 MPa ( 1000 psi) delivery pressure, reducing the speed from 1000 to 500 rpm reduced the measured minimum and maximum clearances by 1 and 3 p.m, respectively (Fig. 12). Increasing the speed from 1000 to 1500 rpm had the reverse effect. The only noticable effect was a general tendency for the amount of cavitation to increase with pump speed.

The delivery pressure was critical in determining the behaviour of the plates. Increasing delivery pressureincreases the plate tilt in all pumps. There appeared to be a substantial hysteris effect in the response of the plate to changes in delivery pressure with the plate attitude being dependent not only on the present delivery pressure, but also on the pressures at which the pump had previously run.

The direction of the plate tilt differed in pump 3 from that found in pumps I and 2. The gear shafts in pomps I and 2 were sealed, producing substantial pressures at the plate centre, whereas in pump 3 the gear shaft region was connected to the pump inlet. It would appear that in all three pomps the direction of the tilt tended to minimize the pump leakage for the particular sealing arrangement at the shaft that had been adopted.

4. Conclusions

As a result of the experimental investigation into the performance of floating wear plates in gear pomps the following conclusions can be drawn: I. The three pumps examined had hydrostatically unbal-

anced wear plates and at high delivery pressures operated with the plates heavily tilted. At lower delivery pressures seal preload tended to stabilize plate operation.

192 E. K W, C.Z Hooke / Wear 209 (1997) 184-192

_9 ~ pumps appear to operate with a high plate tilt, producing a minimum clearance in either the low- or high- pressure region. The plates appear to be tilted so that they occupy the total available end clearances in the unit.

3. Where the central pocket is pressurized the minimum clearance occurs in the low-pressure region. Where the centre is drained, the minimum clearance occurs in the high-pressure region.

4. As well as tilting in response to increasing delivery pressure, all the plates were found to distort. In pump 2 the main distortion consisted o f local bulging o f the plate both inwards and outwards in the vicinity o f the pump centre line. In the other two pumps the distortion appeared to consist o f a general bending o f the plate so as to increase the clearance in the lowest clearance zone.

References

[ I l R.H. Henke. Internal leakage in gear pumps, Appl. Hydraul. 8 ( 1955 ) 63-66.

[ 2 ] CJ. Hnoke, E. KO~T, End plate b-,dance its gear pumps, Prec. Inst. Mech. Eng. Lon. 198B 0984) 55-60.

[3] E. Ko£:, CJ. Hooke, End lubrication and sealing in gear pumps with fixed end plates, J. Fluid Control 18 (3) (1988) 52-69.

[4] E. Koq, CJ. Hooke, The lubrication and sealing of floating end plates in gear pumps, J. Natl Conf. Fluid Power I (1988) 97-I 15.

[5] E. Ko~, CJ. Hooke, The surface non-flatness effect on the design of bush type bearings in high pressure pumps and motors, Wear 135 (1989) 29-39.

[6l E. Kof, An investigation intothe performance of hydrostatically loaded end-plates in high pressure pumps and motors; movable plate design, Wear 141 (1991) 249-265.

[ 7 ] E. Ko~, Analytical and experimental investigation into the sealing and lubrication mechanisms of the gear ends in pumps. Wear 135 (1989) 79-94.

[8l CJ. Hooke, Y.P. Kaknullis, On-line measurement of film thickness, Inst. Mech. Eng. Conf. on Instrument and Computers for Cost Effective Fluid Power Testing. Institute of Mechanical Engineers. London. 1979. Paper CI28179, pp. 51-59.

[9] E. Kof, CJ. Hooke, K.Y. Li, Slipper balance in axial piston pumps and motors, Trans. ASME J. Tribul. i 14 (1992) 766-772.

[ 10] E. Ko¢, CJ. H~ke. Investigation into the effects of orifice size. offset and overclamp ratio on the lubrication of slipper bearings, Tribol. int. 29 (4) (1996) 299-305.

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