Hydrostatic Journal Bearing

SEMINAR REPORT ON INTRODUCTION TO HYDROSTATIC JOURNAL BEARING WITH COAXIAL FLOATING SLEEVE A NEW CONCEPT

Submitted by :-

Guided by :-

# WHY NEED TO APPLY THIS NEW CONCEPT ? (a) Today market demand - high productivity, - high power and high speed range m/c tools. (b) Current available speed only 40-50 k-rpm at 40kw power but some manufacture ask for 100 k-rpm. (c) We can not achieve 100 k-rpm with - ball bearing - conventional hydrostatic bearing

# TYPES OF BEARING: (1) ROLLING CONTACT BEARING (a) Basic four elements – inner and outer race, rolling element like ball and roller, cage. (b) Classification. (c) Antifriction bearing. (d) It is more noisy, low resistance to shock, wear of the balls, not suitable for high load and high rpm. (2) HYDRODYNAMIC BEARING (a) It having thick film between journal and bearing. (b) Load carrying by pressure without any actual contact between journal and bearing. (c) Used for high speed and low load.

(3) HYDROSTATIC BEARING (a) Load is supported by fluid film (b) Lubricant supply under the pressure – externally pressurized bearing. (c) Advantage – high load, even low speed. no starting friction. no rubbing action. (d) Application – used in turbo generator and centrifuge. (4)FLOATING BEARING (a) Floating sleeve located between the journal and bearing surface. (b) Operated at wide range speed for shaft. (c) Used in turbines and compressors.

(5) SQUEEZE FILM JOURNAL BEARING (a) Suitable for oscillate or rotate slowly instruments. (b) Sufficient capacity to carry dynamic load.

Difference between conventional bearing and full-floating bearing

# MAGIC OF COAXIAL FLOATING SLEEVE: (1)

For high speed/power use –In bearing putting sleeve between shaft and housing so splitting total relative speed in two contribution.

(2) Hydrostatic bearing keep all advantage – contactless, no wear phenomenon, no maintenence required. Source of friction is only viscosity of lifting fluid. (3) Sleeve saving half of the friction power because of relation for friction power and relative speed is quadratic.

(4) Because of sleeve improved damping performance because of reduce the stiffness characteristics. (5) Disadvantage of high cost compensate by -High productivity, -Less maintenance, -New market opportunities, due to high speed capacity. (6) We have to focus on radial expansion when working with 100 k-rpm. Because of neglecting centrifugal deformation prevent to reach at request performance

# THE RIGID MODEL:

The model is based on the following hypothesis: (a) the viscosity of the lifting fluid does not depend on temperature, (b) fluid viscosity has a constant value inside each of the two clearances, (c) there is no slip of the fluid at the walls,

(d) the lifting fluid is uncompressible; corrections on this hypothesis are required when using air or gas. (e) laminar flow: velocity profile across the clearances is linear and shear stress is constant. (f) local effects of curvature on fluid flows are neglected as the gap height is negligible with respect to its radius. (g) all eccentricities e are zero. If torque T1 and T2 acting on opposite surface of floating elements are force equal So by intergrating the viscous shear stress τ over the whole bearing friction areas S: S = NP.Af = NP.(AP –3/4AR) T1 =∫ S1 R1.τ 1.dS, T2 =∫ S2 R2.τ 2.dS,

Np = no. of pads Af = pad friction area Ap = area of single pad Ar = area of recess

According to hypothesis (e), considering the viscous shear stress in COUETTE FLOW τ = ( υ µ) / h = (µ∆ ω .R)/h So the expressions of the viscous torques can be rewritten as follows: T1= (µ1 (R1)2 S1 (ω 1 - ω 2))/h1 T2= (µ2 ( R2)2 S2 ω 2)/h2 Now, take T1 = T2, the angular speed ratio ω 1/ω 2 can be written as (ω 1 / ω 2) = 1 + [(µ2 / µ1) (Af2 /Af1) (h1/h2) (R1/R2)2] Since the shaft speed is the independent parameter, speed ratio is expressed as Ω = ω 2/ω 1.

# EXPERIMENTAL RESULT: (1)

Experimental data available for hydrostatic bearing with floating coaxial bush:



Shaft supported by two bearing having 30mm dia. and 4 pads, and external bearing are 39mm dia and 4 pads.



Height of h1 and h2 are 17µ m and 32µ m, working fluid 5cts oil, pressure is 5.5 MPa



Tested up to 50000 rpm.

# THE DEFORMABLE BEARING: (1) In high speed range bearing, due centrifugal force produce the radial expansion which cannot be neglected.

(2)

In figure we can see the FEM analysis.

(3) Radial expansion for any speed obtain from ∆ R - ω diagram and that convert in parabolic polynomial use to be input to matlab program. (4) Calculation of speed ratio is not linear problem because of ω 2 depends on clearance height. unlinear problem solve by step integration if ω 1(t) and ω 2(t) time step, ω 1 is angular speed then next time step (t + ∆ t) ω 2(t + ∆ t)= ω 2(t) + [(T1(t)-T2(t)) ∆ t]/J From the graph Angular speed – time. We see behavior of rigid and deformable bearing.

# RESULT AND ANALYSIS: Ideal speed ratio is 0.5, but we cannot achieve. Why? (1) Minor change in height, viscosity and bearing diameter can be change the performance. (2) Viscosity is affected by temperature – as warm oil flows across the sills. For 1 cst oil we get the difference in speed ratio 0.4 to 0.33 (3) Value of ω for rigid model we take constant for all working speed.

(4)

As per hypothesis (g) we take eccentricity zero but in real condition it never become true.

(5)

Clearance height no longer constant. It become a function of angular coordinate h = h(θ ).

(6) From the graph, Speed ratio- shaft speed we get the performance under different conditions. From bottom 1. Experimental. 2. Low tolerance limit. 3. Numerical deformable. 4. Numerical rigid. 5. High tolerance limit.

# FUTURE DEVELOPMENT AND CONCLUSION: (1) Main consult with height of clearance (h) and viscosity (µ ) for further design improvement in full floating hydro-static bearing. (2) All ball bearing replaced with this one – but very careful finishing and closed tolerance is required. (3) Floating sleeve used only when rotating speed becomes nearly half the one of the shaft means speed ratio (Ω ) is equal to 0.5 (4) Result from proposed method shown good accordance with experimental data.

# REFREENCES: (1) A Textbook of Machine Design by R.S.Khurmi and J. K. Gupta (S. CHAND) (2) Design of Machine Elements by V.B.Bhandari (TATA McGRAW HILL) (3) Hydraulics and Fluid Michanics by Dr.P.N.Modi and Dr.S.M.Seth. A standard book of house (4) Ambrosoni, L., Poli, M., 2002. Theory and experimental tests on a free-floating journal bearing. In: 3rd AIMETA International Tribology Conference, Vietri, Italy. (5) Bassani, R., Piccigallo, B., 1992. Hydrostatic Lubrication. Elsevier, London. (6) Powell, J.W., 1970. Design of Aerostatic Bearings. The Machinery Publishing Co.Ltd., London. (7) Luca ambrosoni, Mario poli, Insitute of Industrial Technology and Automation,Italian National Research Council,Viale Lombradia,201A,20131 Milano,Italy

THANK YOU