Experimental study of hydrodynamic pressure ...

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ScienceDirect Materials Today: Proceedings 2 (2015) 3453 – 3462

4th International Conference on Materials Processing and Characterization

Experimental study of hydrodynamic pressure distribution in oil lubricated two-axial groove journal bearing K.G. Binua, K. Yathisha*, R. Mallyab, B.S. Shenoyc, D.S. Raob, R. Paib a

Department of Mechanical Engineering, St Joseph Engineering College, Vamanjoor, Mangaluru 575028, Karnataka, India b Department of Mechanical & Mfg. Engineering, Manipal Institute of Technology, Manipal 576104, Karnataka, India c Department of Aeronautical & Automobile Engineering, Manipal Institute of Technology, Manipal 576104, Karnataka, India

Abstract Hydrodynamic pressures in oil lubricated two-axial groove finite journal bearing is measured experimentally using a newly developed test rig. Details of design and fabrication of the test rig is presented. Hydrodynamic oil film pressures were measured using two techniques, viz. pressure measurement with sensor mounted on shaft surface and pressure measurement with sensor located outside the bearing assembly. Pressure measurement with first approach was found to yield only the groove pressures. The maximum pressures obtained using second approach is compared with theoretical values. Maximum experimental pressures are found to have a 20% difference in comparison with theoretical maximum pressures. 2014Elsevier The Authors. Ltd. All rights reserved. ©©2015 Ltd. AllElsevier rights reserved. Selection peer-review under responsibility of theof conference committee membersmembers of the 4thofInternational conference conference on Materialson Selectionand and peer-review under responsibility the conference committee the 4th International Processing Characterization. Materials and Processing and Characterization. Keywords: Journal bearing; Hydrodynamic pressure; Lubrication; Pressure sensor;

1. Introduction With ever increasing speed and load on modern day machineries, improved bearing systems are essential to reduce energy losses due to friction and increasing the operational life of machineries. Continuing research in tribology attempts to address this issue by developing lubricants with improved tribological properties and also by developing

* Corresponding author. Tel.: +91-824-2263753; fax: +91-824-2263751. E-mail address: [email protected]

2214-7853 © 2015 Elsevier Ltd. All rights reserved. Selection and peer-review under responsibility of the conference committee members of the 4th International conference on Materials Processing and Characterization. doi:10.1016/j.matpr.2015.07.321

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K.G. Binu et al. / Materials Today: Proceedings 2 (2015) 3453 – 3462

improved bearing configurations. Many theoretical studies have demonstrated the improvement in load carrying capacity of journal bearings operating on nanoparticle lubricant additives [1,2]. Shenoy et al. [3] has theoretically studied the combined influence of nanoparticle lubricant additives and modified externally adjustable bearing configuration on hydrodynamic pressures and the results reveal an improvement in load carrying capacity. These theoretical claims have to be experimentally validated to facilitate industrial implementation. Therefore, an essential requirement in the study of hydrodynamic lubrication is a journal bearing test rig in which performance characteristics of newly formulated lubricants and bearing configurations could be studied. Journal bearing test rig can also be used to gain insights into condition monitoring techniques such as ferrography leading to better selection of lubricants [4]. Key performance parameters in hydrodynamic lubrication include hydrodynamic pressures at the bearing mid-plane, friction force, side leakage, and temperature distribution within the bearing surface. It is also necessary to minimize the pressure losses involved in leading the pressurized oil to the sensor surface. The test rig should also offer flexibility in testing various configurations of bearings. In conventional journal bearing test rigs, hydrodynamic pressures were measured by leading the pressurized oil from the bearing mid-plane through circumferential tap holes to the sensor surface. Christea et al. [5] studied the pressure distribution on a lightly loaded circumferential bearing using a test rig that used pressure tap holes and tubes to deliver pressurized oil to 18 wet-type-strain-gauge pressure sensors. This approach involved pressure loss due to the flow of oil through tap holes and piping leading to the sensor. The presence of air bubbles within the tubing and oil leakage at tap points are other drawbacks of this approach. Similar approach of pressure measurement was also followed by Costa el al. [6], Brito et al. [7-9], and Bouyer and Fillon [10,11] in the study of axial groove journal bearings. Studies have also demonstrated other methods of measuring hydrodynamic pressure. Valkonen A. et al [12] measured the hydrodynamic pressures in a journal bearing using optical pressure sensors which were mounted radially in the bearing assembly. This approach was found to be effective in reducing losses and also demonstrated a novel way of oil supply through the hollow shaft and a radial outlet into the bearing mid-plane. Cabrera et al. [13] used a different approach in measuring hydrodynamic pressures in a water lubricated bearing, wherein a single sub-miniature diaphragm type pressure transducer was mounted on the shaft. The pressurized oil acts on the transducer surface through a radial hole at the bearing mid-plane. The lead wires from the transducer were taken out for data acquisition through the hollow shaft. This eliminates the use of multiple sensors and simplifies the measurement process. Nan Wang et al. [14] used a wireless measurement technique in which six radial holes were drilled at different axial locations on the shaft, each leading to axial guiding holes. The water used as lubricant flows through these holes and acts on the pressure sensors mounted at the end of the shaft. A wireless data acquisition system was used to collect the pressure readings. In this study, a journal bearing test rig is developed to measure the hydrodynamic pressures within a two-axial groove journal bearing. Two approaches of pressure measurement are tried out. In the first approach, a flush type diaphragm pressure sensor is mounted on the surface of a hollow shaft, in line with the bearing mid-plane. Lead wires are taken out through the hollow shaft to the data acquisition system. In the second conventional approach, radial tap holes in the bearing surface are used to transmit pressurized oil from the bearing area to the sensor through lead tubes. A finite bearing with axial-grooves designed in accordance with ESDU [15] standards is used in the study. Initial pressure readings obtained from the study are presented. The maximum pressures are compared with standard theoretical pressures provided in Raimondi and Boyd charts [16]. 2. Details of developed journal bearing test rig The main components of the test rig are: drive unit, test bearing assembly, closed loop oil supply unit, loading unit, pressure measurement system, and data acquisition system. A schematic representation of the complete journal bearing test rig highlighting the components involved is shown in Fig. 1. Details of the sub-assemblies and components involved are described below. 2.1. Drive unit The drive unit consists of a 5 HP AC Motor with speed control unit coupled to a hollow shaft using belt drive. The drive has a maximum speed of 1200 rpm. The hollow mild steel shaft machined is supported on two plummer blocks.


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Fig. 2 provides the shaft dimensions. The bearing area of the shaft was machined using cylindrical grinding to obtain the desired bearing clearance.

Fig. 1. Schematic of journal bearing test rig. 1. Test bearing 2. Test bearing housing 3. Hollow shaft 4. Bearing cover 5. Support bearing 6. Pressure sensor and cap 7. Rotary electric connector 8. Connector holder 9. Amplifier 10. Data acquisition board 11. Computer 12. Downward Weight pan 13. Upward weight pan 14. Friction force scale 15. Pulley 16. Support frame 17. Motor shaft 18. Lubricating oil pressure gauge 19. Oil pump 20. Motor 21. Oil pipe 22. Needle bearing 23. Rubber pad

Fig. 2. Design of shaft

2.2. Test bearing assembly The test bearing assembly consists of Babbitt-lined mild steel two-axial groove test bearing, needle bearing to provide for frictionless rotation of test bearing during operation, and bearing housing. The finite test bearing has an inside diameter and length of 50 mm. The groove dimensions for the test bearing are calculated in accordance with the ESDU data [15]. According to ESDU data [15], the ratio of axial groove length to bearing length a/b= 0.8. Therefore, the length of the axial groove is obtained as a = b × 0.8 = 50 × 0.8 = 40 mm. The circumferential width of the bearing is prescribed as w = 0.25d, (angular extent of about 30°) where, d is the diameter of the bearing. Therefore, width of the groove w = 0.25 × 50 = 12.5 mm. ESDU [15] suggests that the depth of the groove should not be less than 20 times the diametric clearance Cd. Considering a diametric clearance of 0.15 mm (150 microns), depth of the

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groove h should be greater than or equal to 20 × 0.2 = 4 mm. The groove depth was therefore chosen to be 5 mm considering ease of machining. Fig. 3 presents the details of the test bearing along with groove dimensions.

Fig. 3. Test bearing a) Sectional front view at bearing mid-plane b) Sectional side view, and c) Sectional top view

The test bearing is press fitted on to the inner race of a caged needle bearing to facilitate frictionless rotation of the bearing during operation. The needle bearing in turn is held within the bearing housing. The bearing housing assembly is freely mounted on the shaft. The bearing housing is also provided with features to hold four displacement sensors. Oil outlet taps are provided in the bearing housing to collect end leakage from the bearing. Fig. 4 provides details of the bearing housing. 2.3. Closed loop oil feed system The closed loop oil feed unit in the test rig includes a gear pump with filter, pressure regulation valves, end leakage collection tray, and supply tubes. The oil pump used is a Willy Vogel gear pump with 3 liter reservoir capacity. The pump has an in-built oil filter to screen the debris. The feeding pressure was regulated by two precision restrictor valves and monitored by a Bourdon pressure gauge, located at the delivery side of gear pump. The pumped oil enters the bearing mid-plane through two 5 mm circular supply holes provided on the right side face of the bearing leading to the axial-grooves as shown in fig. 3(c). A rectangular oil collector was placed under the bearing to collect the end leakage of oil from the bearing and it was connected to oil sump through supply pipes. This ensured the continuous circulation of oil. The rig thus allowed the regulation of rotational speed, applied load and oil supply pressure. 2.4. Loading unit The loading unit in the test rig is designed to apply a net load on the bearing. The loading system provides for applying load in both vertically upward and downward directions as shown in fig. 5.

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``` Fig. 4. a) Components of bearing housing assembly b) Bearing housing assembly

Fig. 5. Loading unit

2.5. Pressure measurement systems The test rig was designed to facilitate measurement of hydrodynamic pressures using two techniques. In the first approach, pressure sensor was mounted on the shaft surface using a sensor-cap arrangement threaded on the shaft at the bearing mid-plane as shown in fig. 6. The pressurized oil is expected to act on the sensor surface during operation. The variation in hydrodynamic pressure experienced by the sensor during a complete rotation of the shaft is recorded. The lead wires from the sensor are taken out through the hollow shaft as shown in fig. 6.

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Fig. 6. Hydrodynamic pressure measurement using sensor-in-the-shaft approach

The sensor used for the study is a flush type diaphragm pressure transducer. The pressure transducer is of Honeywell make with maximum pressure limit of 20 bar (300 psi). An inline amplifier was connected between the sensor and the data acquisition board to amplify the output signal from the sensor. The leads of the sensor were connected to amplifier through Mercotac electrical connector. The schematic of the measurement process is shown in fig. 7.

Fig. 7. Hydrodynamic pressure measurement process

In the second approach, the pressurized oil was taken out from the bearing mid-plane through tap holes on the side face of the test bearing and sent to the pressure sensor as shown in fig. 8. Fig. 9 provides a developed view of the bearing area showing the location of the tap holes used to measure the hydrodynamic pressures. 12 tap holes are provided on the bearing area such that maximum holes are present on the positive pressure zone. The flush type diaphragm sensor was used at each pressure tap alternatively to obtain the hydrodynamic pressure distribution. 2.6. Data acquisition system As illustrated in fig. 7 the last section of the measurement process is the data acquisition system comprising of a Quanser Q8 data acquisition board and associated QUARC data acquisition software. Using the built-in elements of QUARC software the sensor circuit was built. The QUARC software interfaces with MATLAB to provide the output which could be monitored continuously during experimentation or could be saved as a data file for future processing.


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Fig. 8. Pressure measurement with sensor mounted outside the bearing assembly

Fig. 9. Locations of pressure tap holes across the bearing area

3. Results and discussions The developed journal bearing test rig as illustrated in fig.1 is used to obtain hydrodynamic pressure distribution in a two-axial groove journal bearing using the two pressure measurement techniques illustrated in section 2.5. SAE 30 engine oil was used as the lubricant. Test bearing details, operating conditions and sensor specifications employed in the study are specified in table 1. Initially, an attempt was made to measure hydrodynamic pressures by mounting the pressure sensor on the shaft surface as illustrated in fig. 6. The obtained pressure distribution plot is shown in fig. 10. The plot exhibits two peak pressures of 0.31 MPa. When oil supply to one of the grooves is reduced it causes one of the exhibited peak pressure to also decrease accordingly. This leads us to believe that the sensor was only able to read the oil pressures on both grooves and was not able to sense the hydrodynamic pressures within the land region. A possible reason for this behavior could be the inefficiency of the sensor in reading dynamic pressures. Considering journal speed of 18 rps, the sensor will have to measure hydrodynamic pressures in the circular oil film in 1/18 of a second. The flush type sensor employed has a rated capability of recording 500 readings per second. This would mean that 27 readings could be recorded in the one revolution. As shown in fig. 10, these 27 readings exhibit two peaks which are found to vary with oil supply to the grooves. At very low speeds it was visually evident that pressure spikes were seen only when the sensor approaches the grooves. This leads us to believe that the sensor was not able to read the pressure variation in the 1/18th of a second time duration. A more sensitive sensor might resolve the issue. The approach of having the sensor in the shaft makes the measurement process simple, however requirement of sensor with high dynamic sensitivity might make the process economically challenging.

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Table 1. Operating conditions and sensor specifications Parameters

Values

Shaft diameter D

49.75 mm

Bearing length L

50 mm

Radial clearance C

75 microns

Load range W

450, 850, and 1250 N

Shaft speed N

18 rps

Lubricant viscosity μ

0.1078 Pa.s @ 40°C 0.0137 Pa.s @ 80°C

Pressure sensor Make

Honeywell Model G CP

Range

0 – 300 psi

Accuracy

1% full scale

Fig. 10. Pressure distribution plot using sensor-in-shaft approach

The pressures were then measured using the second approach wherein, the sensor was mounted outside the bearing assembly as explained in section 2.5. Pressurized oil was taken out to the sensor as shown in fig. 8. The journal speed was kept constant at 18 rotations per second and the bearing was subjected to three different loads of 450 N, 850 N and 1250 N. The obtained pressures were validated by comparing them with theoretical pressures obtained using Raimondi and Boyd charts [16]. The theoretical pressures could be read from the charts for specific values of Sommerfeld number. Sommerfeld numbers for hydrodynamic bearings are calculated from the equation given below. 2

S

§ r · PN ¨ ¸ ©C¹ P

(1)

Where, P is the unit pressure obtained as load per unit projected area. P

W LD

(2)

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Sommerfeld numbers are calculated for the three loads, specified in table 1, and the corresponding ratio of ratio of unit pressure P to maximum theoretical pressure Ptmax is read from Raimondi and Boyd chart [16]. The values are shown in table 2. For comparison, table 2 also provides the ratio of unit pressure to experimental maximum pressures Pmax. The theoretical maximum pressures are also computed and compared with experimental maximum pressures. Table 2. Comparison of theoretical and experimental values Load (N)

Sommerfeld Number

P / Ptmax [16]

P / Pmax Experimental

Theoretical maximum pressure Ptmax MPa

Experimental maximum pressure Pmax MPa

450

1.19

0.54

0.64

0.33

0.28

850

0.63

0.53

0.60

0.64

0.56

1250

0.43

0.52

0.59

0.96

0.84

Results reveal that the experimental pressures are lower than the predicted theoretical maximum pressures. A maximum difference of ~20% is observed. This difference in values is most likely due to losses in the loading system. The loading system employed currently is dead weights applied through wires wound around the bearing housing. A hydraulic loading unit with load cell might provide a closer pressure reading. Fig. 11 provides the experimental pressure distribution at a journal speed of 18 rps for three different loads acting on the bearing. The obtained hydrodynamic pressures are in agreement to the theoretical pressures thus validating the experimentation. However, the measurement process needs to be further improved to narrow the difference in experimental and theoretical findings.

Fig. 11. Pressure distribution with the sensor mounted outside the bearing assembly at a journal speed of 18 rps. a) Pressure profile for 450 N load b) pressure profile for 850 N and c) pressure profile at 1250 N

4. Conclusion The developed journal bearing test rig offers hydrodynamic pressures in a two-axial groove bearing in good agreement to theoretical pressures. The sensor-in-shaft approach of measuring hydrodynamic pressures failed to work with the likely reason of the sensor not having the capability of measuring dynamic pressure variations. The

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experimentally obtained maximum pressures using the conventional approach are found to be ~20% lower than the theoretical pressures. Acknowledgements First and second authors would like to acknowledge The Management, St Joseph Engineering College – Vamanjoor, Mangaluru for supporting their research. References [1] K.G. Binu, B.S. Shenoy, D.S. Rao, R. Pai. A variable viscosity approach for the evaluation of load carrying capacity of oil lubricated journal bearing with TiO2 nanoparticles as lubricant additives. Procedia Materials Science 6 (2014) 1051-1067. [2] K.G. Binu, B.S. Shenoy, D.S. Rao, R. Pai. Static characteristics of a fluid film bearing with TiO2 based nanolubricant using the modified Krieger-Dougherty viscosity model and couple stress model. Tribol. Int. 75 (2014) 69-79. [3] B.S. Shenoy, K.G. Binu, R. Pai, D.S. Rao, R.S. Pai. Effect of nanoparticle additives on the performance of an externally adjustable fluid film bearing. Tribol. Int.45(2012) 38 – 42. [4] A. Kiran kumar, P. Archana, K. Rajanarender Reddy. Condition Monitoring of Steam Turbine through Ferrography. IJAMMC 02(2013) doi: http://dx.doi.org/10.11127/ijammc.2013.02.032 [5] A. Cristea, J. Bouyer, M. Fillon, M.D. Pascovici. Pressure and temperature field measurements of a lightly loaded circumferential groove journal bearing, Tribol. T, 54(2011) 806-823. [6] L. Costa, M. Fillon, A.S. Miranda, J.C.P. Claro. Temperature, flow, and eccentricity measurements in a journal bearing with a single axial groove at 90° to the load line. J Tribol-T ASME, 122(1999) 227-232 [7] F.P. Brito, A.S. Miranda, J.C.P. Claro, M. Fillon. Experimental comparison of the performance of a journal bearing with a single and a twin axial groove configuration. Tribol. Int. 54(2012) 1-8. [8] F.P. Brito, J. Bouyer, M. Fillon, A.S. Miranda. Experimental investigation of the influence of supply temperature and supply pressure on the performance of a two axial groove hydrodynamic journal bearing. J Tribol-T ASME, 129(2007) 98-105. [9] F.P. Brito, J. Bouyer, M. Fillon, A.S. Miranda. Thermal behavior and performance characteristics of a twin axial groove journal bearing as a function of applied load and rotational speed. Tribologia-Finnish Journal of Tribology, 3(2006) 24-33 [10] J. Bouyer, M. Fillon. An Experimental Analysis of Misalignment Effects on Hydrodynamic Plain Journal Bearing Performances. J Tribol-T ASME, 124(2002) 313-319. [11] J. Bouyer, M. Fillon. Experimental Measurement of the Friction Torque on Hydrodynamic Plain Journal Bearings during Start-Up.Tribol. Int. 44(2011) pp 772-781. [12] A. Valkonen, J. Juhanko, P. Kuosmanen. Measurement of oil fim pressure in Hydrodynamic Journal Bearing. Presented at the 7th International DAAAM baltic Conference - Industrial Engineering, Tallin Estonia, 2010. [13] D.L. Cabrera, N.H. Woolley, D.R. Allanson,Y.D. Tridimas. Film pressure distribution in water-lubricated rubber journal bearings. Proc. ImechE Part J: J. Engineering Tribology, 219(2005) 125-132. [14] N. Wang, Q. Meng, P. Wang, T. Geng, X. Yuan. Experimental Research on Film Pressure Distribution of Water-Lubricated Rubber Bearing with Multiaxial Grooves. J Fluid Eng-T ASME, 135(2013) 1-6 [15] Engineering Sciences Data Unit (IHS ESDU). Item No. 84031. Calculation methods for steadily loaded axial groove hydrodynamic journal bearings, (1984) [16] A.A. Raimondi, J. Boyd. A solution for the Finite Bearing and its Application to Analysis and Design. ASLE Trans., 1(1958) p 194