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Design and Performance Evaluation of a Miniature Integrated Single Stage Centrifugal Compressor and Permanent Magnet Synchronous Motor Presented by Dipjyoti.

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Presentation on theme: "Design and Performance Evaluation of a Miniature Integrated Single Stage Centrifugal Compressor and Permanent Magnet Synchronous Motor Presented by Dipjyoti."— Presentation transcript:

1 Design and Performance Evaluation of a Miniature Integrated Single Stage Centrifugal Compressor and Permanent Magnet Synchronous Motor Presented by Dipjyoti Acharya M.S.M.E Thesis Defense May 30, 2006, 11: 00 Dept. of Mechanical, Materials and Aerospace Engineering University of Central Florida, Orlando

2 Project Requirements Reverse Turbo Brayton Cycle Cryocooler Single Stage Centrifugal CompressorPermanent Magnet Synchronous Motor Specifications Working fluidAir Stages1 Compression ratio (total to total) 1.58 Inlet condition1 atm, 300K Impeller rev. speed108,000 RPM Mass flow rate7.3 g/s Efficiency0.65 Output Power2000 W Operational Speed200,000 rpm Efficiency @ 77 K89.9 % RotorTitanium StatorSlot less Laminated MagnetSamarium Cobalt WireMulti-strand Litz Wire Winding6 turns/phase/pole

3 Single Stage Centrifugal Compressor

4 Components of the Centrifugal Compressor Impeller Diffuser Inlet Guide Vane Top plateCollector Koford Electric Motor Controller Coupler Courtesy : Ray Zhou, Kevin Finney

5 Compressor Experimental Setup P1, T1P2, T2 P3, T3 P4, T4

6 Experimental Initial Results

7 Issues to be resolved Resolve misalignment issues Design of new coupler to handle high speeds Design a new electric motor to handle the power required by compressor

8 Problems with Rimitec Coupler Servo-insert coupling allowed a restricted amount of misalignment Elastic material in the middle began to plastically deform Stainless steel retaining ring was slip fit Coupler more rigid and prevented it from handling any misalignment Disc couplings style –WM Berg Coupler

9 Coupler Selection Guide Coupler TypeBacklash Allo wed Torque Transmit tal Mis-alignment allowedSpeed Capab ility Overall ra ti ng AxialAngularParallelTorsional Importance Rating Scale2345617 RigidNoneLargeNone Large30 FlexibleJawMedium Small NoneSmall32 Servo-insert/ Disc NoneMedium SmallMedium51 Gear and Spline SmallLargeMedium Small 38 Helical and bello ws SmallLargeMediumSmall 49 Universal joint Medium SmallLargeNoneSmall 37

10 Design of the Helical Coupler 1)Why use a flexible shaft coupling ? 2) Helical Flexure 3) Multiple Starts 4) Flexure Creation Process 5) Material  Martensitic stainless steel CC455 H900 per AMS 5617  Wire EDM, Dynamically balanced while in production  Best for high speed application  Double Start, Ansys Analysis Heli-cal Inc. Rigid couplings would have been always used if it were possible to perfectly align these shafts

11 Helical Coupler Courtesy: Heli-cal Products Company Inc.

12 Alignment Issues and translational stages 1)Straight Edge 2)Laser Alignment Systems - Laser systems 3) Dial indicator methods are A)RIM FACE METHOD B)REVERSE RIM METHOD

13 Reverse Rim Alignment Method SCHEMATIC DIAGRAM OF COMPRESSOR – MOTOR SYSTEM FOR REVERSE RIM ALIGNMENT METHOD  M = the offset in the plane of the movable indicator.  S = the offset in the plane of the stationary indicator.  A = the distance between the stationary and movable dial indicator plungers.  B = the distance from the movable dial indicator plunger to the movable machine’s front feet bolt center.  C= the distance between the movable machines’ front and rear feet bolt centers.

14 Verification of Components With FARO ARM

15 Experimental Results – Translational Stages

16 Experiment Helical Coupler –Fixed Base

17 Compressor Characteristics Curve

18 Available Drivers for Centrifugal Compressor Required speed  108,000 rpm. Achieved Speed  90,000 rpm Electric Motor Air-TurbineTurbocharger Speeds 158,000 rpm250,000 rpm200,000 rpm Accessories ControllerTC-Control System Mist Lubricator Speed Controlling Unit Coupling Flexible Coupler Spline Shaft Cost $ 2500$ 30,000$35,000

19 Design of Permanent Magnet Synchronous Motor

20 Material Selection for Shaft

21 Shear Stress Analysis Maximum Shear Stress Theory Power being transmitted = P = 2000 W Speed of rotation of the shaft = N = 200000 rpm Angular velocity of rotation = ω = 2* π *N/60 Torque developed = T = P/ ω = 0.095 N.m Maximum allowable shear stress = τ max Factor of Safety = F = 3 a = ASME factor for shaft design for shear = ¾ No vertical shear stress Rigidity modulus of the shaft m/l = γ = 117 GPa Angle of shaft twist because of torsion = α (maximum for the outer layer) For a τmax value of 20305 psi (for a Titanium), d = 15.846 mm Thickness of the hollow shaft = t = [D – d]/2 = 0.077 mm << 0.5 mm So, the shaft would not fail under pure shear. Also, Angle of Twist 

22 Bending Stress The values considered for the bending stress are as follows, Elastic modulus of shaft = 116,000 MPa Poisson’s ratio for shaft = υ = 0.34 Ultimate tensile strength of the shaft = 220 MPa Elastic modulus of permanent magnet = 150,000 MPa Poisson’s ratio for permanent magnet(ν) = 0.3 Ultimate tensile strength of the permanent magnet = 82.7 MPa Bending moment due to impeller weight = M = 3.74 N-mm Shaft cross-section – Hollow Shaft Permanent magnet cross-section – Solid Shaft Density of the shaft m/l (ρshaft) = 4500 kg/m3 Density of the permanent magnet (ρmagnet) = 7500 kg/m3 Maximum bending stress (σmax) = M/Z, where Z = section modulus. σmax,Titanium = 0.037 MPa << Ultimate tensile strength of shaft. σmax, Permanent Magnet = 0.048 MPa << Ultimate tensile strength of permanent magnet. So the shaft would not fail under pure bending.

23 Fracture Toughness Fracture Toughness is an issue at cryogenic temperatures. It is defined k = cs(√ π*c)α, where ‘k’ is the ‘Critical Stress Intensity’, Cs – Critical Stress, c – crack length, α – geometry factor (depends on the cross-section of the member) ‘k’ depends on the Bending Stress developed.

24 Centrifugal Stress Analysis Centrifugal stress developed in shaft at 200K rpm = 728 MPa. Centrifugal stress developed in magnet at 200K rpm = 261.2 MPa

25 Thermal Analysis Thermal stress developed in the shaft at 77 K = 329 MPa Thermal stress developed in the magnet at 77 K = 130 MPa Stress due to centrifugal force in shaft rotating at 200,000 rpm = 728 MPa Stress due to centrifugal force in magnet rotating at 200,000 rpm = 251.2 MPa Thermal stress developed in shaft due to operating at 77 K = 329 MPa The Total Stress = 1308.2 MPa < Titanium Grade Yield Strength 1420 MPa So the titanium shaft would not fail. Also, Thermal Stress developed in magnet = 130 MPa < Compressive Strength = 833 MPa. So, the magnet would not crack or crumble to powder.

26 Rotordynamic Analysis of Rotor Courtesy: Dr. Nagraj Arakere, UF, Gainsville

27 Assembly of the Rotor

28 Fabrication of PMSM

29 Results PMSM

30 Motor-Generator Set

31 Design of the Integrated Compressor –Motor System Elimination of the coupler Reduction of number of bearings in the system Usage of fewer components on the rotor to increase the stiffness

32 Rotordynamic Analysis of Integrated System


34 Fabrication of Integrated Rotor

35 Test Rig Structure

36 Features

37 Test Accessories – Electrical Code ComposerEmulator, DSP and Motor ControllerLow Pass FilterPower Meter Courtesy: Liping Zheng and Limei Zhou

38 Test Accessories – Flow P-Transducers Mass Flow Meter

39 Bearing mounting, fit and pre-load Roarke’s Handbook of Stress and Strain Courtesy : Krishna

40 Alignment of Bearings

41 Integrated Compressor – Motor Test Set-up Motor Jacket Stator inside Gas enclosure with adjustable IGV to control tip clearance

42 Two Piece Rotor Free spin test results were successful only to 42,000 rpm Wobbling near the aluminum impeller plug by 0.008 inches Hair crack was visible at the joint Externally Threaded Shaft Internally Threaded Impeller Fabricated Two Piece Rotor

43 Compressor Performance Chart


45 Problems and Future Work in Test Setup Vibrations experienced around 34000 – 37000 rpm range. - leading to stoppage Bearings cooling method to be determined.

46 Conclusion Initial Compressor-Motor Test Setup developed and tested -Helical Coupler was designed and tested -Alignment method was improved by Reverse Rim Method and Translational Stages -Components were verified with FARO Arm and re-fabricated 2 KW Permanent Magnet Synchronous Motor designed and tested -Shaft Material Selection, Stress Analysis Performed and Optimized by Rotordynamic Analysis. Bearing Selection -Fabrication, Assembly Performed and Tested -Motor-Generator set developed to determine motor performance Integrated Compressor – Motor Structure Designed and Tested - Versions of Integrated Rotor was designed and tested -Bearing Fit determined, Pre-load structure designed -Innovative procedure for alignment developed -Adjustable IGV developed for control over tip clearance

47 Thank you

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