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Air Cooled Bearings for Use with High Speed Electric Motors
Lubricants and Lubrication I D. Dowson et al. (Editors) 1995 Elsevier Science B.V.
445
Air Cooled Bearings for Use with High Speed Electric Motors J.E.L. Simmonsa and D. Homerb
a Department of Mechanical Engineering, Heriot-Watt University, Edinburgh, UK. V ickers plc-Michell Bearings, Newcastle upon Tyne, UK. All hydrodynamic bearings require a continuous supply of cooled lubricant. In the case of large, high speed electric motors for industrial applications, oil is most commonly provided by an external lubrication system design to API or equivalent specifications. This paper describes the design and some of the experimental work done in the development of a new range of self contained, air cooled bearings designed to eliminate the need for such costly external lubrication systems. Experimental work was carried out using a 140 mm shaft diameter test bearing operating at a range of duties up to a maximum specific load of 2.07 MPa and a maximum speed of 4500 rev/min. Measurements made include those for oil delivery against speed utilising an innovative, dual mode oil ring described in the paper. 1. INTRODUCTION
Hydrodynamic bearings are fitted to high speed, horizontal shaft, electric motors for a very wide range of industrial applications in preference to rolling element bearings for reasons of longer life and greater reliability. In every application, satisfactory hydrodynamic bearing performance requires ensuring an adequate system for oil supply to the working faces and managing the disposal of heat generated in the bearing. Typical smaller and lower speed machines use an oil ring, which may hang loosely on the shaft or be fixed to it, to generate oil circulation within the bearing, and fins on the casing to dissipate the heat generated. The effectiveness of the oil ring and the limitation of casing surface area are obstacles to extending the simple oil ring concept to larger and faster machinery. For example, difficulties due to high ring speeds and a reduction in oil delivery rate start to appear in two pole motors at shaft diameters above 100 mm. The problem may well be exacerbated if the surface area of the casing is restricted by a centre flange mounting. In previous designs it has been usual when these oil delivery and casing surface area limits have been encountered to employ a completely separate oil
conditioning system to provide cooled lubricant to the bearing. Oil conditioning units, combining the functions of oil circulation and cooling, are generally provided in accordance with API specification 614 which implies the design of complex systems to ensure a reliable supply of oil. The purpose of this paper is to describe the design and underpinning experimental work behind the development of a new range of completely self-contained, air cooled bearings which extend the envelope of machinery size and speed which can be air cooled without the costly overhead of external oil conditioning systems. The internal circulation system for the new bearings makes use of an innovative dual mode oil ring in which oil is collected both from the outside of the ring at start up and run down and from inside the rim of the oil ring at operating speed. Oil collected in this latter fashion is conveyed, under self generated pressure, to an external air blast cooler before being returned to the bearing, still under pressure and in quantities equivalent to those provided by a conventional lubrication system. Figure 1 shows a schematic diagram to illustrate the principle of the dual mode oil ring. Details of its embodiment in a practical bearing design are given in the next section.
446
CONVENTIONAL
SCRAPER\ DUAL MODE OIL
OIL SCOOP
Figure 1. Schematic diagram showing dual mode oil ring 2. BEARING DESIGN
The main features of the test bearing are shown in Figure 2 with basic dimensional and operating duty information in Table 1. The overall size is comparable to that of bearings currently used for medium and large electric motors with a similar conservative loading of the white metal surfaces. The operating speed however is far beyond that of a conventional air cooled design. In the experiments described in the following sections it was demonstrated that the design load and speed combinations given in Table 1 are in fact well within the safe operating scope of the new bearing. The primary duty for this class of bearing is to support radial load, normally the weight of a machine's rotor. This load is carried by a white metal (babbitt) lined journal bush supported in a spherical seating to allow accurate alignment between bush and shaft on setting up. An important secondary requirement in some cases is to provide a location capability in either axial direction. The units which require this feature are equipped with plain thrust faces, machined on the end of the bush, which act against corresponding collars on the shaft.
The bearing casing is split on the horizontal centre line for assembly purposes and designed with a flange mounting for attachment to an electric motor or other high speed machinery. Lubricant is kept from leaking from the bearing, even at high speeds, by non-contacting aluminium labyrinth baffles. The baffles are mounted on the ends of the casing and can be adjusted to suit shaft position. Oil outlet and inlet ports, shown on the lower half casing, are used for pipework connections to a small external oiUair heat exchanger which may be positioned to suit the cooling system of the overall machine. In the case of an electric motor it is usual to position the heat exchanger close to the motor's cooling fan. Table 1 Experimental Bearing: Dimensional Details and Design Duty Journal diameter length diametric clearance
140 mm 85 mm 0.24 mm
Axial location: outside diameter inside diameter surface area
185 mni 145 mm 9335 mm*
Oil ring outside diameter:
400 mm
Casing dimensions: flange diameter axial length
620 mm 290mm
Design duty shaft speed journal load
3600 rev/min 16.4 kN
Lubricant
IS0 VG 32 turbine oil
Cooling air
40°C at 10 m/s
Heat Exchanger
Serck BE1 1 25
447 OIL RING
n
JOURNAL
I OUTLET PORT
I
CONNECTIONS TO EXTERNAL HEAT EXCHANGER
FLAP VALVE
I
INLET PORT
Figure 2. Test Bearing
The dual mode oil ring is mounted on the shaft adjacent to one of the location collars. Operation of the oil ring in supplying lubricant to the working faces and driving oil around the cooling circuit of the bearing is as follows. In the standing condition, prior to shaft rotation, the oil reservoir in the lower hdf of the casing, the heat exchanger and connecting pipe work are filled with oil. At start up oil is picked up on the external periphery of the oil ring and immediately conveyed to the top of the bearing where it is deflected by a scraper built into the interior of the upper half casing. The lubricant flows through ports in the upper casing to supply the annulus on the exterior surface of the bush and thence to oil ways leading to the working surface in a conventional fashion. Simultaneously, some oil from the reservoir adheres to the inner periphery of the oil ring. As speed increases a proportion of this oil on the inner surface is collected by a specially designed scoop positioned in the lower half casing. Conversion from dynamic head to pressure head at the scoop provides the impetus for driving lubricant around the cooling circuit without the need for any additional moving parts. Oil emerging from the outlet port passes via pipework connections to the heat exchanger and back to the inlet port shown on the illustration. From this point the oil flows directly
to the annulus behind the bush from which the working surfaces are supplied. The effect is not only to provide the bearing with a continuous source of cooled oil but also to close the important flap valves fitted to the oil ports in the upper half casing thus pressurising the feed to the bush. A second result of closing the flap valves is the retention of a volume of oil in the upper half casing which is available to flow through the bush during shut down as the shaft comes to rest. In normal, steady state operation oil which has passed through the working surface of the bush drains back to the lower half casing where it is gathered on the inner surface of the oil ring and collected by the oil scoop to complete the cycle and provide a continuous supply for the pressurised lubrication circuit. The effect of the oil collection and circulation system described here is to cause the reservoir of standing oil the lower half casing to empty quickly as the bearing picks up speed so that in normal operation the external periphery of the ring is rotating in air rather than in a bath of oil. In this way the penalty of excessive energy consumption which is characteristic of high speed fixed oil ring systems is avoided.
448 3. IIXI'ERLMENTAI, APPARA'IUS
Motive power for the apparatus is provided by a vatiiible speed DC motor also bolted to the cormnon frame and connected to the shaft via a timing bell drive to achieve the required speed range. 'I'he oil inlet and outlet ports of the test bearing were connected by rigid pipe work to a Serck compact air blasl lieat exchanger. In order i o simulate air conditions in an electriciIl machine adjacent to its cooling fan, the heat exchanger was located in it short lcnglli of ducting about two metres away from Uie bearing. For the majority of the experimenkd work air at 40°C was blown through Lhe duct at a speed of 10 m/s.
'I'he experimental set up, shown in outline form by Figure 3, involved three bearings witli their casings all bolted to i1 rigid franie and linked by ij common shaft. A cenlrally positioned loading bearing, wliose working element consisted of a half bush sitting on top of the shaft, w l z ~ positioned between two support bcarings. 1:xternal load was applied by means of an hydrilulic piston built in bchind Llie half bush of the loading bearing causing it to bear down on tlie shaft. The shaft in its turn reacted against the two supporting heiuitigs, one of which was the test bearing iis described in die previous section. Figure 4 shows an end view of tlie test bearing mounted on the apparatus described here.
VARIABLE SPEED DC MOTOR
\
I-
LOADING BEARING
I
TEST BEARING
J
r------l 1
Figure 3. Diagram of experimental rig
449
Figure 4. Photograph of bearing mounted on experimental rig 4. INSTRUMENTATION
5. EXPERIMENTAL PROGRAMME
Temperature infonnation was monitored throughout the experiments at points in and adjacent to the bearing system using a series of thermocouples connected to a chart recorder. Bush teniperature was measured using two thermocouples embedded 10 mm below the white metal face at bottom dead centre and at 30° from bottom dead centre in the direction of rotation. Further thermocouples were placed in the bearing's upper and lower reservoirs. Oil temperature in the pipe work leading to and from tlie aidoil radiator was similarly monitored as was air temperature on both sides of the radiator. Oil flow rate in the cooling circuit was measured using a turbine flow meter situated in the line close to the inlet to the bearing.
The programme of experiments was arranged to test the bearing at a range of duties up to and in excess of its design duty with the purpose of determining its suitability for typical applications. The tests were not restricted to continuous running but included normal conditions such as start up, run down, over speed and starling in a cold environment. Additional experiments were also carried out to cover emergency and abnormal conditions such as operating with a significantly elevated cooling air temperature and reverse rotation. For the initial series of experiments Uie bearing was operated at its design radial loading and with excess radial loads which were respectively 25% and SO% greater than the design duty. The experiments
450 were carried out over a speed range from 750rev/min to 4500 rev/min. Figure 5 shows measurements of bush, oil bath and cooling air temperature plotted against speed for the most heavily loaded case. The equivalent temperatures for all three loads recorded at synchronous motor speed (3600 rev/min) are given in Table 2. A graph of oil flow rate at inlet lo the bearing plotted against speed up to 4500 rev/min for the 50% overload condition is shown by Figure 6. It can be seen that oil flow, which is evidently more than adequate to maintain bearing temperature within working limits, increases linearly with rotational speed to reach an amount in excess of 1300 Vh at the top speed of 4500 rev/min . Table 2 Bush and Oil Bath Temperatures at Synchronous Speed (3600 rev/min): Standard Duty and Overload Conditions
Journal Load Pressure (kN) (MPa)
Temperature (“C) Bush Bath Cooling Air
16.4 20.5 24.6
79.4 81.7 83.5
1.38 1.72 2.07
57.4 59.1 59.0
42.7 42.8 42.8
In and bearing temperatures were attained. particular there was no delay in generating cooling oil circulation at start up. A typical set of resultant temperatures is given in Table 3. The bearing’s tolerance to conditions which were likely to lead to higher operating temperatures was investigated in two ways. Firstly, in one experiment the cooling air temperature was raised to 60°C while maintaining the normal air flow rate of 10 m/s. In a second experiment, cooling air velocity was restricted to 5 m/s which led in turn to the air temperature also increasing. One might imagine circumstances such as these two cases arising when there is a need to keep vital equipment running during an emergency or if service management and control equipment is in some way defective. The results of the elevated temperature tests are shown in Table 3. Bush temperatures in the region of 100°C as recorded here are well within operating limits for white metal lined bushes and the experiments were only terminated when typical alarm and shut down temperatures were reached, not for any concern for the safety or integrity of the bearing. Table 3 Bush and Oil Bath Temperatures at Synchronous Speed (3600 rev/min): Abnormal Conditions
Journal In addition to the tests at design duty, over speed and overload conditions shown by Figure 5 and Table 2, a series of experiments were carried out to test the bearing system’s resilience to temporary or abnormal conditions which may be experienced during service life. One important condition which might be expected to occur on a regular basis is that of starting up in a cold environment. Start up and operation in very low temperature surroundings were simulated by substituting the normally used IS0 VG 32 turbine oil with higher viscosity IS0 VG 100 oil. This latter oil has a viscosity at room temperature comparable with that of 1SOVG32 oil in the range O°C to 5OC. No problems were encountered in tests starting from cold and continuing until steady operating conditions
Load
Temperature
(“0
(kN)
Bush
Bath
Cooling Air
Comment
16.4
87.6
59.4
41.9
24.6
95.1
74.3
63.7
16.4
100.3
85.9
57.8
Simulated cold environment Elevated air temperature Reducedair flow
Clearly, the bearing’s oil scoop is designed for a preferred direction of rotation. Nevertheless occasional reverse rotation may be a planned event in some applications such as when a pump is driven
45 1 100
*
90
--
80
--
70
-4
60 * -
+ +
50 *.
40
-.
30
..
0
H
A
A
1500
2000
m u A
500
+
+ A
+
4
4
I
Oil Bath
H
H
A
A
A
A Cooling Air
A
1000
+
Journal Bush
2500 3OOO Shaft Speed (rev / mln)
3500
4000
4500
5000
Figure 5. Bearing bush temperature and oil bath temperature plotted against speed for an applied load of 2.07 MPa L
1400
H
-
z5
1200
H
I
L
.m
1000
1
0)
t
0
-
6
800
H
600
.n 400 1500
H
2000
2500
3000 3500 Shall Speed (rev / rnin)
4000
Figure 6. Oil flow rate plotted against speed for applied load of 2.07 MPa
4500
5000
452 temporarily as a turbine as part of a shut down sequence. Alternatively, reverse rotation is conceivable as an accidental condition resulting from a fault elsewhere in the system such as incorrect wiring of an electric motor. In either circumstance it is necessary for the bearing to be capable of reverse rotation operation for a reasonable period. Reverse running experiments for the bearing loaded to its design duty demonstrated that it could continue to operate perfectly satisfactorily at 3600 rev/min for 30minutes before a nominal bearing alarm temperature of 9O0C was reached. 6. DISCUSSION
The results of the experimental programme as indicated by Figure 5 and by Table 2, show that the subject bearing operates satisfactorily over a very wide range of speeds and loads. In this case the bearing was tested for extended periods at loads 50% greater than the required standard design duty and over a wide speed range which extended 25% beyond synchronous motor speed. Oil flow rate, shown plotted against speed in Figure 6, is more than adequate for a bearing of this size and is a good testament to the soundness of oil circulation system design. Thus, Figure 6 shows a linear increase in oil flow with increasing speed up to and well beyond the working range of conventional oil ring systems. There has been little prior work published showing delivery rates for fixed oil rings which is perhaps surprising given their wide use in many plain bearings. Such work as has been published (1, 2) is concerned with conventional rings where the oil is collected only from the external periphery of the ring using a scraper and at much lower speeds than apply in this case. It is well known in practice is that at high speeds the effectiveness of conventional rings is limited by a reduction in oil delivery due to the effects of centripetal acceleration on oil initially adhering to the ring. Hence, Gardner (21, working with oil rings suitable for marine propulsion bearings reports oil delivery more than halved at the top end of his speed range. Simmons and Advani (3) suggest an accepted maximum peripheral speed for a conventional fixed oil ring of about 33 m/s. This compares with 94 m/s
which is the peripheral speed of the oil ring at 4500 rev/min in the experiments reported here. In addition to the design, over speed and overload duties, the bearing system was exposed to a series of abnormal conditions likely to have caused distress to a conventional, air cooled bearing. The results of these experiments are summarised by Table 3. No apparent problems arose during the experimental work, nor were the ultimate limits of perfomlance for white metal bearings approached. Similarly the bearing was operated in reverse for a satisfactory period without a n y ill effects. All this evidence points to a robust design capable of sustained operation in demanding conditions. The bearing described in this paper was designed to accommodate significant radial loads associated with electrical motors and other high speed machinery. As indicated earlier provision is made for notional axial loading for location purposes by means of the plain thrust faces on the ends of the bush and the matching collars on the shaft. Clearly the bearing as described is capable of development to allow for much more substantial axial loading through the incorporation of tilting pad thrust faces on the bush and correspondingly larger thrust collars on the shaft. 7. CONCLUSION This paper has introduced the design and experimental work behind the introduction of a new design of self-contained bearing for use in highspeed machinery including synchronous electric motors. The bearing incorporates a dual mode fixed oil circulation ring in which the oil is collected in conventional fashion from the outside of the ring at start up and from the inside of the peripheral rim at higher speeds. Sufficient pressure is generated by this internal pump to drive the gathered lubricant through an external air cooled heat exchanger and back to the bearing as a cooled supply. The design has proved satisfactory for a wide range of conditions which extend significantly the envelope of duties that can be satisfied by a bearing system relying solely on air cooling. This advance has been achieved with little in the way of additional cost or complexity when compared with a conventional ring
453 oiled bearing. The general robusmess of the new design has been confirmed in a series of experiments in which the bearing has performed well in the face of a number of exceptional operating conditions. 8. ACKNOWLEDGEMENTS
The authors are grateful to the Directors of Michell Bearings for permission to publish material included in this paper, to Mr G. Humble for his assistance in the experimental work and to other colleagues at Michell Bearings who contributed to this project.
REFERENCES 1. D.C. Lemmon and E.R. Booser, 'Bearing oil-ring performance', ASME Paper 59-Lub-5. 2. W.W. Gardner, 'Bearing oil delivery by diskscraper means', ASME Paper 76-Lub-2, Prw. Joint Lubrication Conference, Boston, Mass., 5-7 October 1976. 3. J.E.L. Simmons and S.D. Advani, 'The shrouded ring bearing - an advance for large, high performance, self-contained systems', Roc. IMechE Seminar on Developments in Plain Bearings for the '90s'. 17 May, 1990. pp 55-63.
445
Air Cooled Bearings for Use with High Speed Electric Motors J.E.L. Simmonsa and D. Homerb
a Department of Mechanical Engineering, Heriot-Watt University, Edinburgh, UK. V ickers plc-Michell Bearings, Newcastle upon Tyne, UK. All hydrodynamic bearings require a continuous supply of cooled lubricant. In the case of large, high speed electric motors for industrial applications, oil is most commonly provided by an external lubrication system design to API or equivalent specifications. This paper describes the design and some of the experimental work done in the development of a new range of self contained, air cooled bearings designed to eliminate the need for such costly external lubrication systems. Experimental work was carried out using a 140 mm shaft diameter test bearing operating at a range of duties up to a maximum specific load of 2.07 MPa and a maximum speed of 4500 rev/min. Measurements made include those for oil delivery against speed utilising an innovative, dual mode oil ring described in the paper. 1. INTRODUCTION
Hydrodynamic bearings are fitted to high speed, horizontal shaft, electric motors for a very wide range of industrial applications in preference to rolling element bearings for reasons of longer life and greater reliability. In every application, satisfactory hydrodynamic bearing performance requires ensuring an adequate system for oil supply to the working faces and managing the disposal of heat generated in the bearing. Typical smaller and lower speed machines use an oil ring, which may hang loosely on the shaft or be fixed to it, to generate oil circulation within the bearing, and fins on the casing to dissipate the heat generated. The effectiveness of the oil ring and the limitation of casing surface area are obstacles to extending the simple oil ring concept to larger and faster machinery. For example, difficulties due to high ring speeds and a reduction in oil delivery rate start to appear in two pole motors at shaft diameters above 100 mm. The problem may well be exacerbated if the surface area of the casing is restricted by a centre flange mounting. In previous designs it has been usual when these oil delivery and casing surface area limits have been encountered to employ a completely separate oil
conditioning system to provide cooled lubricant to the bearing. Oil conditioning units, combining the functions of oil circulation and cooling, are generally provided in accordance with API specification 614 which implies the design of complex systems to ensure a reliable supply of oil. The purpose of this paper is to describe the design and underpinning experimental work behind the development of a new range of completely self-contained, air cooled bearings which extend the envelope of machinery size and speed which can be air cooled without the costly overhead of external oil conditioning systems. The internal circulation system for the new bearings makes use of an innovative dual mode oil ring in which oil is collected both from the outside of the ring at start up and run down and from inside the rim of the oil ring at operating speed. Oil collected in this latter fashion is conveyed, under self generated pressure, to an external air blast cooler before being returned to the bearing, still under pressure and in quantities equivalent to those provided by a conventional lubrication system. Figure 1 shows a schematic diagram to illustrate the principle of the dual mode oil ring. Details of its embodiment in a practical bearing design are given in the next section.
446
CONVENTIONAL
SCRAPER\ DUAL MODE OIL
OIL SCOOP
Figure 1. Schematic diagram showing dual mode oil ring 2. BEARING DESIGN
The main features of the test bearing are shown in Figure 2 with basic dimensional and operating duty information in Table 1. The overall size is comparable to that of bearings currently used for medium and large electric motors with a similar conservative loading of the white metal surfaces. The operating speed however is far beyond that of a conventional air cooled design. In the experiments described in the following sections it was demonstrated that the design load and speed combinations given in Table 1 are in fact well within the safe operating scope of the new bearing. The primary duty for this class of bearing is to support radial load, normally the weight of a machine's rotor. This load is carried by a white metal (babbitt) lined journal bush supported in a spherical seating to allow accurate alignment between bush and shaft on setting up. An important secondary requirement in some cases is to provide a location capability in either axial direction. The units which require this feature are equipped with plain thrust faces, machined on the end of the bush, which act against corresponding collars on the shaft.
The bearing casing is split on the horizontal centre line for assembly purposes and designed with a flange mounting for attachment to an electric motor or other high speed machinery. Lubricant is kept from leaking from the bearing, even at high speeds, by non-contacting aluminium labyrinth baffles. The baffles are mounted on the ends of the casing and can be adjusted to suit shaft position. Oil outlet and inlet ports, shown on the lower half casing, are used for pipework connections to a small external oiUair heat exchanger which may be positioned to suit the cooling system of the overall machine. In the case of an electric motor it is usual to position the heat exchanger close to the motor's cooling fan. Table 1 Experimental Bearing: Dimensional Details and Design Duty Journal diameter length diametric clearance
140 mm 85 mm 0.24 mm
Axial location: outside diameter inside diameter surface area
185 mni 145 mm 9335 mm*
Oil ring outside diameter:
400 mm
Casing dimensions: flange diameter axial length
620 mm 290mm
Design duty shaft speed journal load
3600 rev/min 16.4 kN
Lubricant
IS0 VG 32 turbine oil
Cooling air
40°C at 10 m/s
Heat Exchanger
Serck BE1 1 25
447 OIL RING
n
JOURNAL
I OUTLET PORT
I
CONNECTIONS TO EXTERNAL HEAT EXCHANGER
FLAP VALVE
I
INLET PORT
Figure 2. Test Bearing
The dual mode oil ring is mounted on the shaft adjacent to one of the location collars. Operation of the oil ring in supplying lubricant to the working faces and driving oil around the cooling circuit of the bearing is as follows. In the standing condition, prior to shaft rotation, the oil reservoir in the lower hdf of the casing, the heat exchanger and connecting pipe work are filled with oil. At start up oil is picked up on the external periphery of the oil ring and immediately conveyed to the top of the bearing where it is deflected by a scraper built into the interior of the upper half casing. The lubricant flows through ports in the upper casing to supply the annulus on the exterior surface of the bush and thence to oil ways leading to the working surface in a conventional fashion. Simultaneously, some oil from the reservoir adheres to the inner periphery of the oil ring. As speed increases a proportion of this oil on the inner surface is collected by a specially designed scoop positioned in the lower half casing. Conversion from dynamic head to pressure head at the scoop provides the impetus for driving lubricant around the cooling circuit without the need for any additional moving parts. Oil emerging from the outlet port passes via pipework connections to the heat exchanger and back to the inlet port shown on the illustration. From this point the oil flows directly
to the annulus behind the bush from which the working surfaces are supplied. The effect is not only to provide the bearing with a continuous source of cooled oil but also to close the important flap valves fitted to the oil ports in the upper half casing thus pressurising the feed to the bush. A second result of closing the flap valves is the retention of a volume of oil in the upper half casing which is available to flow through the bush during shut down as the shaft comes to rest. In normal, steady state operation oil which has passed through the working surface of the bush drains back to the lower half casing where it is gathered on the inner surface of the oil ring and collected by the oil scoop to complete the cycle and provide a continuous supply for the pressurised lubrication circuit. The effect of the oil collection and circulation system described here is to cause the reservoir of standing oil the lower half casing to empty quickly as the bearing picks up speed so that in normal operation the external periphery of the ring is rotating in air rather than in a bath of oil. In this way the penalty of excessive energy consumption which is characteristic of high speed fixed oil ring systems is avoided.
448 3. IIXI'ERLMENTAI, APPARA'IUS
Motive power for the apparatus is provided by a vatiiible speed DC motor also bolted to the cormnon frame and connected to the shaft via a timing bell drive to achieve the required speed range. 'I'he oil inlet and outlet ports of the test bearing were connected by rigid pipe work to a Serck compact air blasl lieat exchanger. In order i o simulate air conditions in an electriciIl machine adjacent to its cooling fan, the heat exchanger was located in it short lcnglli of ducting about two metres away from Uie bearing. For the majority of the experimenkd work air at 40°C was blown through Lhe duct at a speed of 10 m/s.
'I'he experimental set up, shown in outline form by Figure 3, involved three bearings witli their casings all bolted to i1 rigid franie and linked by ij common shaft. A cenlrally positioned loading bearing, wliose working element consisted of a half bush sitting on top of the shaft, w l z ~ positioned between two support bcarings. 1:xternal load was applied by means of an hydrilulic piston built in bchind Llie half bush of the loading bearing causing it to bear down on tlie shaft. The shaft in its turn reacted against the two supporting heiuitigs, one of which was the test bearing iis described in die previous section. Figure 4 shows an end view of tlie test bearing mounted on the apparatus described here.
VARIABLE SPEED DC MOTOR
\
I-
LOADING BEARING
I
TEST BEARING
J
r------l 1
Figure 3. Diagram of experimental rig
449
Figure 4. Photograph of bearing mounted on experimental rig 4. INSTRUMENTATION
5. EXPERIMENTAL PROGRAMME
Temperature infonnation was monitored throughout the experiments at points in and adjacent to the bearing system using a series of thermocouples connected to a chart recorder. Bush teniperature was measured using two thermocouples embedded 10 mm below the white metal face at bottom dead centre and at 30° from bottom dead centre in the direction of rotation. Further thermocouples were placed in the bearing's upper and lower reservoirs. Oil temperature in the pipe work leading to and from tlie aidoil radiator was similarly monitored as was air temperature on both sides of the radiator. Oil flow rate in the cooling circuit was measured using a turbine flow meter situated in the line close to the inlet to the bearing.
The programme of experiments was arranged to test the bearing at a range of duties up to and in excess of its design duty with the purpose of determining its suitability for typical applications. The tests were not restricted to continuous running but included normal conditions such as start up, run down, over speed and starling in a cold environment. Additional experiments were also carried out to cover emergency and abnormal conditions such as operating with a significantly elevated cooling air temperature and reverse rotation. For the initial series of experiments Uie bearing was operated at its design radial loading and with excess radial loads which were respectively 25% and SO% greater than the design duty. The experiments
450 were carried out over a speed range from 750rev/min to 4500 rev/min. Figure 5 shows measurements of bush, oil bath and cooling air temperature plotted against speed for the most heavily loaded case. The equivalent temperatures for all three loads recorded at synchronous motor speed (3600 rev/min) are given in Table 2. A graph of oil flow rate at inlet lo the bearing plotted against speed up to 4500 rev/min for the 50% overload condition is shown by Figure 6. It can be seen that oil flow, which is evidently more than adequate to maintain bearing temperature within working limits, increases linearly with rotational speed to reach an amount in excess of 1300 Vh at the top speed of 4500 rev/min . Table 2 Bush and Oil Bath Temperatures at Synchronous Speed (3600 rev/min): Standard Duty and Overload Conditions
Journal Load Pressure (kN) (MPa)
Temperature (“C) Bush Bath Cooling Air
16.4 20.5 24.6
79.4 81.7 83.5
1.38 1.72 2.07
57.4 59.1 59.0
42.7 42.8 42.8
In and bearing temperatures were attained. particular there was no delay in generating cooling oil circulation at start up. A typical set of resultant temperatures is given in Table 3. The bearing’s tolerance to conditions which were likely to lead to higher operating temperatures was investigated in two ways. Firstly, in one experiment the cooling air temperature was raised to 60°C while maintaining the normal air flow rate of 10 m/s. In a second experiment, cooling air velocity was restricted to 5 m/s which led in turn to the air temperature also increasing. One might imagine circumstances such as these two cases arising when there is a need to keep vital equipment running during an emergency or if service management and control equipment is in some way defective. The results of the elevated temperature tests are shown in Table 3. Bush temperatures in the region of 100°C as recorded here are well within operating limits for white metal lined bushes and the experiments were only terminated when typical alarm and shut down temperatures were reached, not for any concern for the safety or integrity of the bearing. Table 3 Bush and Oil Bath Temperatures at Synchronous Speed (3600 rev/min): Abnormal Conditions
Journal In addition to the tests at design duty, over speed and overload conditions shown by Figure 5 and Table 2, a series of experiments were carried out to test the bearing system’s resilience to temporary or abnormal conditions which may be experienced during service life. One important condition which might be expected to occur on a regular basis is that of starting up in a cold environment. Start up and operation in very low temperature surroundings were simulated by substituting the normally used IS0 VG 32 turbine oil with higher viscosity IS0 VG 100 oil. This latter oil has a viscosity at room temperature comparable with that of 1SOVG32 oil in the range O°C to 5OC. No problems were encountered in tests starting from cold and continuing until steady operating conditions
Load
Temperature
(“0
(kN)
Bush
Bath
Cooling Air
Comment
16.4
87.6
59.4
41.9
24.6
95.1
74.3
63.7
16.4
100.3
85.9
57.8
Simulated cold environment Elevated air temperature Reducedair flow
Clearly, the bearing’s oil scoop is designed for a preferred direction of rotation. Nevertheless occasional reverse rotation may be a planned event in some applications such as when a pump is driven
45 1 100
*
90
--
80
--
70
-4
60 * -
+ +
50 *.
40
-.
30
..
0
H
A
A
1500
2000
m u A
500
+
+ A
+
4
4
I
Oil Bath
H
H
A
A
A
A Cooling Air
A
1000
+
Journal Bush
2500 3OOO Shaft Speed (rev / mln)
3500
4000
4500
5000
Figure 5. Bearing bush temperature and oil bath temperature plotted against speed for an applied load of 2.07 MPa L
1400
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z5
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0)
t
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.n 400 1500
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Figure 6. Oil flow rate plotted against speed for applied load of 2.07 MPa
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452 temporarily as a turbine as part of a shut down sequence. Alternatively, reverse rotation is conceivable as an accidental condition resulting from a fault elsewhere in the system such as incorrect wiring of an electric motor. In either circumstance it is necessary for the bearing to be capable of reverse rotation operation for a reasonable period. Reverse running experiments for the bearing loaded to its design duty demonstrated that it could continue to operate perfectly satisfactorily at 3600 rev/min for 30minutes before a nominal bearing alarm temperature of 9O0C was reached. 6. DISCUSSION
The results of the experimental programme as indicated by Figure 5 and by Table 2, show that the subject bearing operates satisfactorily over a very wide range of speeds and loads. In this case the bearing was tested for extended periods at loads 50% greater than the required standard design duty and over a wide speed range which extended 25% beyond synchronous motor speed. Oil flow rate, shown plotted against speed in Figure 6, is more than adequate for a bearing of this size and is a good testament to the soundness of oil circulation system design. Thus, Figure 6 shows a linear increase in oil flow with increasing speed up to and well beyond the working range of conventional oil ring systems. There has been little prior work published showing delivery rates for fixed oil rings which is perhaps surprising given their wide use in many plain bearings. Such work as has been published (1, 2) is concerned with conventional rings where the oil is collected only from the external periphery of the ring using a scraper and at much lower speeds than apply in this case. It is well known in practice is that at high speeds the effectiveness of conventional rings is limited by a reduction in oil delivery due to the effects of centripetal acceleration on oil initially adhering to the ring. Hence, Gardner (21, working with oil rings suitable for marine propulsion bearings reports oil delivery more than halved at the top end of his speed range. Simmons and Advani (3) suggest an accepted maximum peripheral speed for a conventional fixed oil ring of about 33 m/s. This compares with 94 m/s
which is the peripheral speed of the oil ring at 4500 rev/min in the experiments reported here. In addition to the design, over speed and overload duties, the bearing system was exposed to a series of abnormal conditions likely to have caused distress to a conventional, air cooled bearing. The results of these experiments are summarised by Table 3. No apparent problems arose during the experimental work, nor were the ultimate limits of perfomlance for white metal bearings approached. Similarly the bearing was operated in reverse for a satisfactory period without a n y ill effects. All this evidence points to a robust design capable of sustained operation in demanding conditions. The bearing described in this paper was designed to accommodate significant radial loads associated with electrical motors and other high speed machinery. As indicated earlier provision is made for notional axial loading for location purposes by means of the plain thrust faces on the ends of the bush and the matching collars on the shaft. Clearly the bearing as described is capable of development to allow for much more substantial axial loading through the incorporation of tilting pad thrust faces on the bush and correspondingly larger thrust collars on the shaft. 7. CONCLUSION This paper has introduced the design and experimental work behind the introduction of a new design of self-contained bearing for use in highspeed machinery including synchronous electric motors. The bearing incorporates a dual mode fixed oil circulation ring in which the oil is collected in conventional fashion from the outside of the ring at start up and from the inside of the peripheral rim at higher speeds. Sufficient pressure is generated by this internal pump to drive the gathered lubricant through an external air cooled heat exchanger and back to the bearing as a cooled supply. The design has proved satisfactory for a wide range of conditions which extend significantly the envelope of duties that can be satisfied by a bearing system relying solely on air cooling. This advance has been achieved with little in the way of additional cost or complexity when compared with a conventional ring
453 oiled bearing. The general robusmess of the new design has been confirmed in a series of experiments in which the bearing has performed well in the face of a number of exceptional operating conditions. 8. ACKNOWLEDGEMENTS
The authors are grateful to the Directors of Michell Bearings for permission to publish material included in this paper, to Mr G. Humble for his assistance in the experimental work and to other colleagues at Michell Bearings who contributed to this project.
REFERENCES 1. D.C. Lemmon and E.R. Booser, 'Bearing oil-ring performance', ASME Paper 59-Lub-5. 2. W.W. Gardner, 'Bearing oil delivery by diskscraper means', ASME Paper 76-Lub-2, Prw. Joint Lubrication Conference, Boston, Mass., 5-7 October 1976. 3. J.E.L. Simmons and S.D. Advani, 'The shrouded ring bearing - an advance for large, high performance, self-contained systems', Roc. IMechE Seminar on Developments in Plain Bearings for the '90s'. 17 May, 1990. pp 55-63.