DESIGN OF A BUSH PRESSING MACHINE FOR PUMPS

DESIGN OF A BUSH PRESSING

MACHINE FOR PUMPS

AMITH KALEKAR 1

Student, M.E. –Mech. (Design Engg.), Walchand Institute of Technology (Solapur University) Near Ashok Chowk, Solapur, Maharashtra 413006, India

amithkalekar50@gmail.com

S. B. TULJAPURE 2

Assistant Professor - Mech. Dept., Walchand Institute of Technology (Solapur University) Near Ashok Chowk, Solapur, Maharashtra 413006, India

stuljapure@gmail.com

Abstract

Bearings are among the most important components in the vast majority of machines. Bush is an independent plain bearing that is inserted into a housing to provide a bearing surface for rotary applications. Bushes are fitted either by pressing for less interference or by heating the frame or cooling the bushes using dry ice or liquid nitrogen. In the present case, due to uneven cooling using liquid nitrogen or other problems, bush is not fitting easily in some of the pump frames. So there is a need of a bush pressing machine. A ‘C’ frame type hydraulic bush pressing machine is designed in this work.

Keywords: Design of press, Interference pressure, Pressing force, Bush. 1. Introduction

The press machines are basically of two types. One is Mechanical Presses & second is Hydraulic Presses.For over 30 years, there has been a clear trend toward hydraulic presses. Until recently, the most common production press has been the mechanical press- a force delivering machine in which power is derived from a rotated crankshaft. In recent years hydraulic presses surpassed mechanical presses. Following are the advantages of hydraulic presses.

1.1. Advantages of Hydraulic Press machines:

1. Full power stroke

2. Built-in overload protection

3. Much lower original cost and operating costs 4. Larger capacities at lower cost

5. More control flexibility 6. Greater versatility 7. Quiet operation 8. More compact 9. Safety

So it was decided to design a hydraulic press machine for pressing bush into the pump frame.

2. Literature Review

R. Lewis & et al (2005)developed anultrasound as a tool to non-destructively determine press-fit contact pressures. The interface contact pressure has been determined for a number of different press and shrink fit cases. The results showed a central region of approximately uniform pressure with edge stress at the contact sides. The magnitude of the pressure in the central region agreed well with the elastic Lame’s analysis. The average contact pressure in a shrink fit and press fit joint were similar. However in the shrink fit joint more uneven contact pressure was observed with regions of poor conformity. This could be because the action of pressing on a sleeve plastically smoothes out long wavelength roughness leading to a more conforming surface.

H.N.Chauhan & M.P.Bambhania (2013) designed & analysed frame of a 63 tonne power press machine using Finite Element method. Due to the impact loading at the end of the bolster plate there was development of a crack at the corner and stress generated was more due to continuous loading and stress concentration. Modifications were done by introducing the fillets of proper size. Also plate thickness was reduced which saved material.

Bhavesh Khichadia, Dipeshkumar Chauhan (2014) took review on design and analysis of mechanical press frame. They dealt with the analytical method and corresponding design and analysis of mechanical press frame. Modeling of press frame has created by the CAD software and according to the modeling structure failure analysis done by FEA tool. Stress distributions in press frame have been found out by the analytical and simulation methods. Reference data were used for the new design for modification of a new press structure. With regard to design specification, stress distribution, deflection, optimization, ergonomics, stiffness and rigidity was focused on recent design and development in press frame obtained from structural components of press machine frame.

B. Parthiban and et al (2014) designed a ‘C’ type hydraulic press structure and cylinder. They analyzed press frame and cylinder to improve its performance and quality of press working operation. The frame and cylinder were modeled by using software CATIA. Structural analysis was done using analysis software ANSYS. An integrated approach was developed to verify the structural performance and stress strain distributions were plotted by using ANSYS software. According to the structural values the dimensions of the frame and cylinder were modified to perform the functions satisfactory.

D. Ravi (2014) analyzed a power press of 10 tonne capacity under static condition. The modeling of the C- frame power press was done using Pro/E software. The 3D model of the power press was analyzed in static condition to find the stresses and deflections in the structure. Later part involved the reduction in weight of the power press by varying or reducing the thickness of frame and bed and the press was analyzed in static condition to find the results. The result obtained from analysis package is within the limit.

Review of present situation:

After studying the drawings of bush & frame, following points are noted.

Table 1 Dimensions of Bush & Frame

Part Parameters

Inner Diameter (mm) Outer Diameter (mm) Length of bush or

thickness of casing (mm)

Bush 27.5+0.1 36.00 +0.08/+0.065 26.5

Casing/Frame hole 36+0.025 100 26.5

Maximum interference is 80 micron & Minimum interference is 40 micron.

Reduction in diameter of bush after nitrogen cooling and its effect: Various charts related to liquid nitrogen cooling are studied. There is 5 micron reduction per mm with liquid nitrogen at -2000C for bronze. [12].

Table 2 Dimensions of Bush (before & after nitrogen cooling)

Frame Hole Diameter (mm) Bush O.D. Before cooling (mm) Bush OD After cooling (mm)

Difference in Diameter Hole Dia. – Bush Dia. (mm)

36.000 36.025 36.065 36.080 36.022 36.037 36.025- 36.022 = 0.003 mm

36.000-36.037= - 0.037 mm 36.000-36.022 = -0.022 mm 36.025-36.037 = -0.012 mm

Thus it is observed that some cases might exist in which bush outer diameter will be greater than hole diameter. Hence initially there is need of redefining the tolerances. Also from the study of tolerance charts, it is observed that highest interference fit for a bush of diameter 36 mm is 36H7u6 which is of force fit category.

Considering this, the tolerance charts further studied & it is observed that in the interference fit group, there are total four categories.

1. Press fit

3. Drive fit-II

4. Force fit

In the above categories, the amount of interference goes on increasing from Press fit to Force fit. For the present problem, the amount of tolerance for the above categories is

Table 3 Types of interference fits & Tolerances

Sr. No. Category of fit Tolerance on shaft (in microns)

1 Press fit +42, +26

2 Drive fit +50, +34

3 Drive fit +59, +43

4 Force fit +76, +60

From the above data & considering the variation caused during the nitrogen cooling of bush, it was decided to design a bush press machine for maximum tolerance which is obtained from the first category (Press fit).

Table 4 Comparison of inteference between nitrogen cooling & press fit category

Sr. No.

Outer diameter of bush after

Nitrogen Cooling

Outer Diameter of bush considering press fit

(I category)

Frame Hole Diameter

(mm) (mm) (mm)

1 36.022 36.042 36.000

2 36.037 36.026 36.025

From the above table, it is seen that a maximum interference of 42 microns occurs for press fit category for this size of bush.

3. Calculation of interference pressure

(a)

Fig. 2 Front view showing (a) cylinder assembled with an interference fit & (b) hub & hollow shaft disaasembled (also showing interference pressure) [ref.2]

The total radial interference can be calculated as below. [Ref.2]

r = rh - rs = rf pf

 

pf is the pressure caused by interference between the shaft or bush & the hub or flange. r = Total radial interference

rh = Radial displacement of hub/frame rs = Radial displacement of shaft or bush

rf = Outer Radius of bush

ro = Outer Radius of hub/ frame

ri = Inner Radius of bush

Eh = Modulus of elasticity of hub or frame material

Es = Modulus of elasticity of shaft or bush material h = Poisson’s ratio of hub or frame material

s = Poisson’s ratio of shaft or bush material

4. Properties of materials of bush & frame:

Poisson’s ratio of Cast Iron FG260 is 0.26 Modulus of elasticity is 128 GPa Poisson’s ratio of phosphor Bronze is 0.35 Young’s Modulus of phosphor Bronze is 111 GPa

r = rf pf

 

0.021 =18 xpf

.

.

.

.

pf = 26.96 MPa

5. Calculation of pressing force

The force required for press fit F = 2  rf l pf where,

 = Coefficient of friction between the mating parts rf = Outer Radius of bush

l = Length of bush pf = Interference pressure

Here,

Coefficient of friction (Kinetic or sliding frictional coefficient) between Bronze-Cast Iron =  = 0.22 for Clean and Dry Surfaces.

Force required for press fit F = 2  x 0.22 x 18 x 26.5 x 26.96 = 17.776 KN ≈ 17.78 KN

Verification of the above calculations was also done using the programmed calculators available on the internet.

6. Selection of material of press machine body/frame

After studying the applications of various types of materials, finally mild steel with 0.3 to 0.4 % carbon & 0.3 to 0.6 % Manganese is selected for the frame or body of the machine.

Standard designation of this material is C 35

Following are its properties.

Yield Stress = σ = 31 kg/mm2 = 304 N/mm2 Modulus of Elasticity = E = 2.06 x 105 N/mm2 Poisson’s ratio =  = 0.3

7. Design of press machine 7.1. Design of Press Body/Frame

The permissible tensile stress for the plates is given by

σt = = _. = 121.6 N/mm2

Here factory of safety is taken as 2.5 which is sufficient for this case as system is hydraulic and for higher than set pressures, relief valve will open & oil flow will be diverted to reservoir.

Since two identical plates (C shaped plates) are taken, The force on each plate is (17.78/2) = 8.89 KN

These ‘C’ plates are subjected to direct tensile stress & bending stresses. The stresses are maximum at the inner fiber.

At the inner fibre,

σt = x

σt = Permissible tensile stress P = Force acting

A = resisting area of C plate = Width of C plate x Thickness of C plate Mb = Bending moment

y = Distance of neutral fiber from the point of application of force I = Moment of inertia of resisting cross sectional area of plate Let us take, Width of ‘C’ plate = 150 mm

Hence,

121.6 = x

t = 4.87 mm

Let us take a standard 6 mm thick plate for C plates of the body.

7.2. Design of plates for job mounting

The maximum force acting on the plates is 17.78 KN.

Let us take 250 mm x 300 mm (width) as area of mounting plates of job. Now, load acts at the centre of the plate.

So, this is a case of a simply supported beam of length 300 mm with a central load of 17.78 KN. Thus, the plate will be only under bending load.

Using,

σt =

For simply supported beam,

Max. Bending Moment = Mb = (Force x Length )/ 4 = (17780 x 300)/4 = 1333.5 KN-mm

121.6 =

t = 16.22 mm

t = 18 mm (Say)

Now the dimensions of the job mounting plate are 250 mm x 300 mm x 18 mm

The calculations done for job mounting plate are also applicable to hydraulic cylinder mounting plate. This plate will have same dimensions as job mounting plate.

7.3. Frame strengthening plates: Four plates are used for keeping two ‘C’ plate together & strengthening the body of the machine. The dimensions of these plates are 300 mm (width) x 200 mm (height) x 10 mm (thickness). Two plates are mounted in the front side, one at the top side & one at the bottom side. One plate is mounted at the back side of the machine.

Plates for leveling & installation of machine: Four brackets are provided at four corners of the machine. These are to be used for leveling the machine & installing the machine finally.

7.4. Design of hydraulic system

Here meter out circuit is used. In this circuit, the oil returning from the cylinder is controlled. [Ref. mujum]

The choice of this circuit is done due to following characteristics of this circuit.

i. No loss of pressure to the actuator as the full oil pressure is fed to the cylinder. Even at no-load the actuator is subjected to maximum pressure.

ii. The actuator movement is more stable.

iii. Heat generated due to throttling is fed to the oil reservoir.

iv. Suits both over running loads as well as speed control of hydro-motors. As the cylinder is fed with the entire pump pressure, there is possibility of higher friction loss.

Fig. 3 Bush Pressing Machine

8. Selection of hydraulic cylinder, pump & electric motor

Cylinder bore = 80 mm

Calculation of required oil pressure:

Pressure = =  = 3.53 N/mm2 = 3.53 x 106 N/m2 = 35.3 bar

Piston stroke is finalized considering the joub mounting plate & position of cylinder. Piston Stroke = 300 mm

The maximum capacity of cylinder is 140 Kg/cm2 (13.734 N/mm2) which is more than the actual working pressure = 3.53 N/mm2

So the cylinder model selected is safe.

8.1. Check for Safe Stroke against Buckling Failure:

Actual stroke length is 300 mm which is less than 2912.5 mm Hence piston rod is safe under buckling load.

8.2. Selection of pump: Gear pumps are having simple construction & hence they are cheap. So gear pump is selected to give pressurized oil to hydraulic cylinder. This is capable of giving 63 bar pressure continuously which is more than the required 35 bar pressure. It gives a discharge of 75.3 liter per minute at a speed of 1500 rpm is chosen. This discharge is on higher side of our requirement. But it can be operated at a 400 rpm to get lesser discharge. So it will operate at 400 rpm & will deliver 20.08 liters per minute.

8.3. Selection of motor:

Power required to drive pump shaft (P) = Pressure x Discharge

P = 3.53 x 106 x 20.08 x 10-3 / 60 = 1181.37 watts = 1.2 kilowatts (approx.)

Considering efficiency of the motors & referring the selection chart, an electric motor of 1.5 Kilowatt output power & having 8 pole is selected which will give a speed 375 rpm which very near to the requirement.

9. Conclusion

Following are the conclusions.

 The present interference is on higher side. This interference should be reduced to the allowable extent.

 Keeping frame or body of pump in sunrays may help to some extent.

10. References

[1] V. B. Bhandari, “Design of Machine Elements”, Tata McGraw-Hill Publishing Co. Ltd., New Delhi, 2008 [2] Bernard Hamrock, “Fundamentals of Machine Elements”, McGraw-Hill Publication, 1999

[3] R. S. Khurmi, J. K. Gupta “A Text Book of Machine Design”, S. Chand Publications, New Delhi 2012

[4] S. R. Majumdar “Oil Hydraulic Systems-Principles and Maintenance”, Tata Mc-Graw Hill Publishing Co. Ltd. New Delhi, 2008 [5] PSG Design Data Handbook, 2008

[6] R. Lewis, M.B. Marshall, R.S. Dwyer-Joyce “Measurement of Interface Pressure in Interference Fits”, Proceedings of the Institution of Mechanical Engineers, Part C: Journal of Mechanical Engineering Science, 2005, 219(2), 127-139

[7] Thomas A. Cherukara, Glen H. Besterfield, Autar K. Kaw, “Parametric Analysis and Ultimate Testing of Bascule Trunnion Assemblies”, Ninth Biennial Symposium, "Preserving Traditional Values with New Technologies, October 22-25,2002

[8] H.N.Chauhan & M.P.Bambhania, “Design & Analysis of Frame of 63 Ton Power Press Machine by Using Finite Element Method”, Indian Journal of Applied Research, Volume: 3, Issue:7, July 2013

[9] Bhavesh Khichadia, Dipeshkumar Chauhan, “A Review on Design And Analysis of Mechanical Press Frame”, International Journal of Advance Engineering and Research Development, Volume 1, Issue 6, June 2014

[10] B. Parthiban and et al, “Design and Analysis of C type hydraulic press structure and cylinder”, International Journal of Research in Aeronautical and Mechanical Engineering, Vol.2 Issue.3, March 2014.Pgs: 47-56

[11] D. Ravi, “Computer Aided Design and Analysis of Power Press”, Middle-East Journal of Scientific Research 20 (10): 1239-1246, 2014