Design, Development and Testing of Parallel offset Coupling with Angular offset

Design, Development and Testing of Parallel offset Coupling with Angular offset

1Archana Chandak 2Anurag Nema 3Dr. F. B. Sayyad

1PG Student, Design Engineering GSMCOE Balewadi, Pune

2Assist. Professor, dept of Mechanical Engineering DPCOE, Wagholi, Pune

3Professor, dept of Mechanical Engineering GSMCOE Balewadi, Pune

Abstract- Parallel and angular offset Couplings were developed to fill a gap in the family of torque- rigid couplings. Most couplings are designed to accommodate axial, angular, or parallel shaft displacements only. For some applications, however, the operational conditions require all possible shaft misalignments. If these shaft misalignments exceed the limit of the selected coupling capacity, excess side loads are introduced into the equipment which can cause vibrations, life reduction or failure of vital machine components such as bearings, motors, etc.

The Parallel and angular offset Couplings are a modification of the Inline Coupling, designed to accommodate 5 degrees of angular shaft misalignment. This coupling allows easy adjustment to any possible misaligned shaft position without imposing heavy side loads on shafts, bearings or other machine equipment. This Couplings offer large shaft misalignment capabilities and constant angular velocity. The acting forces within the coupling can be precisely calculated, assuring a sound coupling design which is especially important for heavy-duty applications.

Key words: Parallel and angular offset coupling, Misalignment, axial load and Power Transmission.

1. INTRODUCTION

   Shaft misalignment has major implications for modern-day rotating equipment reliability. Although effective alignment techniques have been applied successfully on a wide range of equipment for some time, deterioration of the alignment state can frequently occur due to, for example, changes in equipment operating conditions. Because of this rigid support, it is virtually impossible to avoid slight misalignments between a driving and driven shaft when they are connected. Restoring forces that occur as the two coupled shafts compete to maintain their original positions can put unwanted strain on shaft bearings, causing them to wear out prematurely. Additional axial loads are also placed on the bearings as thermal growth occurs in shafting during operation.

   This situation can lead to the imposition of excessive forces on the equipment rotating and static elements, most commonly resulting in bearing or coupling failure. In extreme circumstances contact between rotating and stationary components can be expected to occur. The presence of shaft misalignment can greatly influence

   machinery vibration response. However, its detection through vibration diagnostics is not a straightforward matter due to the lack of a clear understanding of the physical mechanism relating shaft misalignment to vibration. multi-harmonic response from rotor dynamic systems subjected to angular and parallel misalignment by assuming coupling transmitted forces represented by a half- sinusoid function having fundamental frequency equal to twice rotational speed. Assumptions and investigated the transient response of a misaligned rotor system.

   1.1 PARALLEL OFFSET MISALIGNMENT

   Offset misalignment, sometimes referred to as parallel misalignment, is the distance between the shaft centers of rotation measured at the plane of power transmission. This is typically measured at the coupling center. The units for this measurement are mils (where 1 mil = 0.001 in.). A measure of the offset distance between the centerlines of driving and driven shafts. Coupling catalogs will show the maximum parallel misalignment tolerable in each coupling. A coupling should not be operated with both parallel and angular misalignment at their maximum values.

   Fig 1.1 Parallel Offset

   Fig 1.2 Angular Offset

2. EXPERIMENTAL ANALYSIS

   The Couplings were developed to fill a gap in the family of torque-rigid couplings. Most couplings are designed to accommodate axial, angular, or parallel shaft displacements only. For some applications, however, the operational conditions require all possible shaft misalignments. If these shaft misalignments exceed the limit of the selected coupling capacity, excess side loads are introduced into the equipment which can cause vibrations, life reduction or failure of vital machine components such as bearings, motors, etc. The Couplings are a modification of the Inline Coupling, designed to accommodate 5 degrees of angular shaft misalignment. This coupling allows easy adjustment to any possible misaligned shaft position without imposing heavy side loads on shafts, bearings or other machine equipment. The Couplings offer large shaft misalignment capabilities and constant angular velocity. The acting forces within the coupling can be precisely calculated, assuring a sound coupling design which is especially important for heavy- duty applications. The experimental setup as shown in the figure-2.1.

   Fig-2.1 Experimental Setup

   Table-2.1 Description of parts

   Part no.	Description	Material
   1.	FRAME	MS
   2.	BRG_HSG_L_PLATE	EN9
   3.	BRG_HSG	EN9
   4.	MAIN PULLEY	EN9
   5.	IP_SHAFT	EN24
   6.	OP_SHAFT	EN24
   7.	DRIVER _DISK	EN24
   8.	INT_DISK	EN24
   9.	DRIVEN_DISK	EN24
   10.	LINKS	EN9
   11.	D_D_PINS	EN24
   12.	I_D_PINS	EN24
   13.	SLIDE BAR	EN9
   14.	SLIDE NUT	EN9
   15.	CLAMP PLATE	EN9
   16.	MOTOR PLATE	MS
   17.	HANDLE	MS
   18.	BOLT REST	EN9
   19.	MOTOR	STD
   20.	BELT(6 X 600)	STD
   21.	Bearing 6204ZZ	STD
   22.	Bearing 6203ZZ	STD
   23.	Bearing 6200ZZ	STD
   24.	Grub screw M8 x 8	STD
   25.	Grub screw M6 x 8	STD
   26.	HEX BOLT M8 x 25	STD
   27.	HEX BOLT M10 x 30	STD
   28.	HEX BOLT M10 x 50	STD
   29.	HEX BOLT M10 x 200	STD

3. DESIGN OF COMPONENTS

   1. SELECTION OF DRIVE MOTOR

      The metric system uses kilowatts (kW) for driver ratings.

      T=

      Converting kW to torque: KW ×84518

      RPM

      Where, T = the torque in inch pounds KW= the motor or other kilowatts

      RPM = the operating speed in revolutions per minut 84518 = a constant used when torque is in inch-pounds.

      Use 7043 for foot-pounds, and 9550 for Newton-meters

      0.3 = ×9550

      1200

      = 0.038

      Thus the minimum input power required will be 38 watt. Thus selecting a drive motor as follows

      DRIVE MOTOR

      Type: – Single Phase Ac Motor.

      Power: – 1

      15

      Hp (50 Watts)

      Voltage: – 230 Volts, 50 Hz

      Current: – 0.5 Amps Speed: – Min = 0 Rpm

      Y = 0.484 2.86

      Z

      2.86

      Max = 6000 Rpm

   2. DESIGN OF GEAR DRIVE FROM MOTOR TO

      Yp = 0.484

      24 = 0.058

      2.86

      INPUT SHAFT 3.2.1DESIGN OF SPUR PINION & GEAR

      Fig 3.1 Spur Pinion

      Fig3.2 Spur Gear

      Stage: Drive as gear and pinion arrangement Maximum load =Maximum Torque / Radius of gear Maximum torque = 0.4 N-m

      No of teeth on gear = 120 Module = 1.275mm

      Radius of gear by geometry =120×1.275 = 76.5mm

      2

      Maximum load = T =0.4×103 = 5.3

      Yg = 0.484 120 = 0.460

      Syp = 4.930

      Syg = 39.10

      As Syp< Sys, Pinion is weaker

      WT = (Syp) x b x m

      =4.93 m x m

      WT = 4.93 m2———- (B)

      Equation (A) & (B)

      4.93 m2 =8 m=1.274mm

      Selecting standard module = 1.275 mm, for ease of construction as we go for single stage gear box for making size compact and achieving maximum strength and proper mesh.

      Designation	Ultimate Tensile Strength N/Mm2	Yield Strength N/Mm2
      En 24	800	680

   3. SELECTION OF INPUT SHAFT Table-3.3 Stress Values from Data Book

      76.5

      b = 10 m

      Material of spur gear and pinion = Nylon-66 Sult pinion = Sult gear= 85 N/mm2

      Service factor (Cs) = 1.5

      The gear and pinion arrangement where as pinion has 10 teeth and gear has 30 teeth share the entire tooth load Pt = (W x Cs) =8 N.

      P eff = 8 N (as Cv =1 due to low speed of operation) P eff = 8 N ——– (A)

      Lewis Strength equation WT = S b y m

      Where;

      Fig-3.3 Input shaft

      Fig-3.4 Output shaft

   4. SELECTION OF DRIVER DISK HUB

      Driver disk hub can be considered to be a hollow shaft subjected to torsional load.

      Fig 3.5 Driver Disk

      Fig 3.6 Intermediate Disk

      Fig-3.7 Driven Disk

   5. SELECTION OF LINK

      Fig-3.8 Link

   6. MISCELLANEOUS PARTS DRAWING

   Fig-3.9 Housing of Input Bearing

   Fig 3.10 Housing of Output Bearing

   Fig-3.11 I D PIN

   Fig 3.12 Support Rib

   Fig 3.13 Output shaft

   Fig-3.14 Input Coupler

   Fig-3.15 Coupler Shaft

   Fig-3.16 Output Coupler

4. ANALYSIS OF COMPONENTS

   4.1 ANALYSIS PROCEDURE

   1. Modeling of the geometry is being done in Unigraphics software.

   2. The generated IGES file is exported to ANSYS workbench

   3. The model is discretised into finite elements by triangular mesh elements.

   4. Applying boundary conditions and loads.

   5. Solve the problem.

   1. ANALYSIS OF INPUT SHAFT

      . Fig-4.1 Stress distribution on input shaft

      Stress distribution on input shaft as shown in fig-4.1, as the maximum stress induced in the material (1.17 N/mm2) < allowable stress (144 N/mm2) the input shaft is safe under pure Torsional load.

   2. ANALYSIS OF INPUT COUPLER

      Fig-4.2 Stress distribution on input coupler

      Stress distribution on input coupler as shown in fig-4.2, as the maximum stress induced in the material (1.25 N/mm2)

      < allowable stress (144 N/mm2) the IP Coupler is safe under pure torsional load.

   3. ANALYSIS OF DRIVER DISK

      Fig-4.3 Stress distribution on driver disk

      Stress distribution on driver disk as shown in fig- 4.3, as the maximum stress induced in the material (1.54 N/mm2) < allowable stress (144 N/mm2) the Driver Disk is safe under pure torsional load.

   4. ANALYSIS OF INTERMEDIATE DISK

      Fig-4.4 Stress distribution on intermediate disk

      Stress distribution on intermediate disk as shown in fig-4.4, As the maximum stress induced in the material (1.24 N/mm2) < allowable stress (144 N/mm2) the Intermediate Disk is safe under pure torsional load.

   5. ANALYSIS OF COUPLER SHAFT

      Fig-4.5 Stress distribution on Coupler Shaft

      Stress distribution on coupler Shaft as shown in fig-4.5, As the maximum stress induced in the material (3.39 N/mm2) < allowable stress (144 N/mm2) the Coupler shaft is safe under pure torsional load.

   6. ANALYSIS OF OUTPUT SHAFT

      . Fig-4.6 Stress distribution on output shaft

      Stress distribution on output shaft as shown in fig-4.6, as the maximum stress induced in the material (1.17 N/mm2)

      < allowable stress (144 N/mm2) the input shaft is safe under pure torsional load.

   7. ANALYSIS OF OUTPUT COUPLER

      Fig-4.7 Stress distribution on output coupler

      Stress distribution on output coupler as shown in fig-4.7, as the maximum stress induced in the material (1.16 N/mm2) < allowable stress (144 N/mm2) the OP Couplar is safe under pure torsional load.

   8. ANALYSIS OF DRIVEN DISK

   Fig-4.8 Stress distribution on driven disk

   Stress distribution on driven disk as shown in fig-4.8, as the maximum stress induced in the material (1.54 N/mm2)

   < allowable stress (144 N/mm2 ) the Driven Disk is safe under pure torsional load.

5. RESULTS AND DISSCUSIONS

   1. COMPARATIVE ANALYSIS OF ANGULAR OFFSET PERFORMANCE OF COUPLING

      1. Torque analysis

         Fig-5.1 Variation of Torque V/s Speed for different angular offset angles

         Variation of torque for different angular offset angles is shown in fiig-5.1. It is observed that the torque values remain almost same for all angular offset settings.

      2. Output Power analysis

         Fig-5.2 Variation of Power Output V/s Speed for different angular offset angles

         Variation of power output for different angular offset angles is shown in fig-5.2. It is seen that there is marginal drop in output power with increase in angular offset thus it can be safely stated that coupling offers maximum power output at minimum angular offset.

      3. Efficiency analysis

         Fig-5.3 Variation of Efficiency V/s Speed for different angular offset angles

         Variation of efficiency for different angular offset angles is shown in fig-5.3. It is seen that there is marginal drop in efficiency with increase in angular offset thus it can be safely stated that coupling offers maximum efficiency at minimum angular offset.

   2. COMPARATIVE ANALYSIS OF PARALLEL OFFSET PERFORMANCE OF COUPLING

      1. Torque analysis

         Fig-5.4 Variation of Torque V/s Speed for different parallel offset angles

         Variation of torque for different parallel offset angles is shown in fig-5.4. It is observed that the torque values remain almost same for all parallel offset settings.

      2. Output Power analysis

         Fig-5.5 Variation of Power Output V/s Speed for different parallel offset angles

         Variation of power output for different parallel offset angles is shown in fig-5.5. It is seen that there is a marginal drop in output power with increase in parallel offset thus it can be safely stated that coupling offers maximum power output at minimum parallel offset.

      3. Efficiency analysis

         Fig-5.6 Variation of Efficiency V/s Speed for different parallel offset angles

         Variation of efficiency for different parallel offset angles is shown in fig-5.6. It is seen that there is marginal drop in efficiency with increase in parallel offset thus it can be safely stated that coupling offers maximum efficiency at minimum parallel offset.

6. CONCLUSIONS

From the experimental setup of parallel and angular offset coupling, the following results were obtained

• The maximm displacement or offset in parallel condition is 35mm on either side of mean.

• Torque transmitted by the coupling drops with increase in speed marginally.

• Maximum efficiency of coupling is achieved when operated at zero offset, but there is a marginal decrease in efficiency as offset is increased.

• The coupling can transmit angular offset from 1o to 5o

  .The angular offset is adjustable in step-less manner meaning that even an offset angle of 3.9 o is possible.

• Maximum efficiency of coupling is achieved when operated at zero angular offset, but there is a marginal decrease in efficiency as angular offset is increased.

REFERENCES

1. Irvin Redmond-Saudi Arabian Oil Company, Dhahran 31311, Eastern Province, Saudi Arabia (2013) Shaft Misalignment and Vibration – A Model

2. Redmond -Dynamic Analysis Unit, Saudi Aramco, R- 99,Bldg. 9155, Dhahran 31311, Saudi Arabia (2010) Study of a misaligned flexibly coupled shaft system having nonlinear bearings and cyclic coupling stiffness Theoretical model and analysis.

3. Mr. S.B. Jaiswal Prof. M.D. Pasarkar IJETA (ISSN 2250- 2459, Volume 2, Issue 5, May (2012) Failure Analysis of Flange Coupling In Industry.

4. Why Shaft Misalignment Continues to Befudle and Undermine Even the Best CBM and Pro-Active Maintenance Programs, Proc. Of The Predictive Maintenance Technology National Conference, Indianapolis, In 5 : 18-23, Dec 3-6, 1996.

5. Misalignment As a Source of Vibration in Rotating Shaft Systems, Proc. Intl. Modal Analysis Conf. (IMAC) XIX, Orlando, Feb. 2001.

6. Design of machine elements V. B. Bhandari

7. Theory of machinesS.S.Ratan.

8. Gear coupler Manual by Lovejoy

9. Alignment of vertical shaft hydro units united states department of the interior bureau of reclamation.

10. Shaft Alignment Handbook by John Piotrowski.

11. PSG Design Databook.