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NASA CR-1 74866 

MTI 85TR20 I 

I 



« 



IW\SA 



SEAL TECHNOLOGY FOR 
LIQUID OXYGEN (LOX) TURBOPUMPS | 



by Wilbur Shapiro and Robert Hamm 



MECHANICAL TECHNOLOGY iNCORPORATED 



Prepared for 
NATIONAL AERONAUTICS AND SPACE ADMINISTRATION 



NASA Lewis Research Center 
Contract NAS3-23260 



N88~13603 



(NASA-CR-17<»866) SE^L TECHNOLOGY FOE LIQUID 
nvvrPN 'LOXJ TURBOPUMPS Final Report, aeo, 

388 p SViil: NT15 HC A17/HF M01 ti^i. ^^^^^ 0114097 



1. Rtport No. 

NASA CR-1 74866 



2. Gov«rnm«nt AccMtion No. 



3. Rtcipisnt's Catalog No. 



4. Titia and Subtitia 

Seal Technology for Liquid Oxygen (LOX) Turbopumps 



5. Report Data 

November 1985 



6. Performing Organization Coda 



7. Author(s) 

Wilbur Shapiro 
Robert Haimn 



8. Parformfng Organization Report No. 
85TR20 



10. Worl< Unit No. 



9. Performing Organization Name and Addrui 

Mechanical Technology Incorporated 
Research and Development Division 
968 Albany-Shaker Road 
Latham, New York 12110 



11. Contract or Grant No. 
NAS3-23260 



12. Sponsoring Agency Name and Addreu 

National Aeronautics and Space Administration 
Lewis Research Center 
Cleveland, OH 44135 



13. Type of Report and Period Covered 

FINAL 

2/82 - 11/85 



14. Sponsoring Agency Code 



15. Supplementarv Note* 

Project Manager: J. A. Hemmingir, NASA-Lewis Research Center, Cleveland, OH 44135 



16. Abstract 

Two types of advanced seals for liquid oxygen (LOX) turbopumps were investigated. One 
was a apirai-groove face seal whose function is to seal high-pressure LOX at the 
impeller end of the turbopump. The other was a floating-ring, Rayleigh-step, helium 
buffered seal used to prevent LOX ingress to the turbine side of the unit. For each 
seal type, two sizes were investigated - 50 and 20 mm. A turbine-driven test rig was 
designed and manufactured, and a test program was completed on the SO-tran floating- 
ring, Rayleigh-step, helium buffered seal. Significant results of the progvam were: 

Spiral-Groove LOX Seal 

• Vaporization in the flow path could cause seal failure by overheating; therefore, 
the spiral-groove pumping portion of the seal that provides the fluid film must 
circulate fliud without disruption if vaporization occurs in the sealing dam. This 
is successfully accomplished by a pressure-balanced spiral-groove concept that is 
described in the text. 

• The spiral-groove configuration is affected by turbulence in the fluid film and 
pressure drops due to fluid inertia at sudden contractions. The net. result of these 
effects are deep grooves, large operating films, and high power loss when compared 
against seals operating with laminar films. Turbulence and inertia are induced by 
the high-density and low-viscosity characteristics of LOX. 



17. Key Words (Suggested by Author (s) I 

Rotating Shaft Seals 
Floating-Ring Seals 
Self-Acting Seals 
Turbopump Shaft Seals 



Spiral-Groove Seals 
Liquid-Oxygen Seals 
Helium Buffer Seals 
Rayleigh Step Seals 



18. Distribution Statement 



19. Security Classif. (of this report) 



20. Security Classlf. (of this page) 



21. No. of Pages 



22. Price 



* For sale by the National Technical information Service, Springfield, Virginia 22151 



^^•I-l7983 



NASA-C-lfi8 (Rev. f)-7n 



• Computed flow levels for the 50-mm seal are approximately 30 mm /s (A. 7 gpm) at 
pressure levels of 5.17 MPa (750 psi). Film thickness is 0.0236 mm (0.93 mils) and 
the interface speed is la3 m/s (600 ft/s). 

The floating-ring, Rayleigh-step seals were exposed to an extensive testing program. 
Successful operation was demonstrated at speeds up to 5760 m/s (55,000 r/min), and 
buffer fluid pressure levels of 1379 kPa (200 psia). Average leakage rates per ring 
were 0.0016 kg/s (20.2 scfm). Multiple high acceleration starts at 152 m/s^ (500 
ft/s ) were accomplished without incident. Several seal failures did occur, but were 
attributed to external influences such as excessive rig vibration and contamination of 
the helium supply system. 

The program clearly pointed out the c.eed to consider system environmental factors such 
as thermal and centrifugal distortions and rotor vibrations in the seal design. More 
liberal seal clearances would probably have permitted operation to 7330 rad/s (70,000 
r/min), but rig changes made to improve dynamic response altered the seal environment 
and caused greater than anticipated clearance closure due to thermal growth. 



ACKNOWLEDGMJ_NTS 

The work reported upon herein was due to Che contribution of several key indi- 
viduals at Mechanical Technology Incorporated (MTI). Their efforts, which are 
presented below, are gratefully acknowledged. 

Individual Description 

Dr. Jed Walowit Spiral-Groove Seal Analysis 

Mr. Antonio Artilles Rayleigh-Step Helium Seal Analysis 

Mr. Henry Jones Seal Design 

Mr. John Dunne Test Rig Design 

MTI expresses appreciation to Mr. Jeffrey Frazier of Wyle Laboratories, 
Norco, California, for his efforts during the testing phase of the program. 
In addition, MTI expresses sincere gratitude to Messrs. Joseph Hemminger and 
J. P. Wanhainen of NASA/LeRC for their consistent support and assistance 
throughout the course of the program. MTI also acknowledges, posthumously, 
the efforts of Mr. Lawrence Ludwig of NASA/LeRC who was responsible for 
advanced seal development and conceived ttis seal configurations discussed in 
this report. 



-Ill- 



I 

I 

E 



I. 



TABLE OF CONTENTS 

SKTION PAGE 

ACKNOWLEDGMENTS ii 

LIST OF FIGURES viii 

LIST OF TABLES 

1.0 INTRODUCTION 1-1 

2.0 SUMMARY OF SIGNIFICANT RESULTS AND CONCLUSIONS 2-1 

2.1 Configuration and Principle of Operation of a 50-mni Helium 

Buffer Seal 2-1 

2.2 Test Results of the 50-mm Rayleigh-Step Helium Buffer Seal . 2-2 

2.3 Conclusions and Recommendations for the 50-mm 

Rayleigh-Step Seal 2-4 

2.4 Results and Conclusions for the 20-ttim Rayleigh-Step Helium 

Buffer Seal 2-7 

2.5 Description and Principle of Operation of the Spiral-Groove 

LOX Seals , 2-8 

2.5.1 Special Considerations for LOX 2-8 

2.5.2 Pressure-Balanced Spiral-Groove Seal 2-10 

2.6 50-mm Spiral-Groove Seal 2-11 

2.6.1 Groove Geometry 2-11 

2.6.2 Summary of Pe: '^ormance 2-12 

2.7 20-mra Spiral-Groove L.-al 2-13 

2.7.1 Geometry 2-13 

2.7.2 Summary of Performance 2-13 

2.8 Test Rig 2-14 

2.8.1 Design Philosophy 2-14 

2.8.2 General Configuration 2-14 

2.8.3 Summary of Test Rig Performance 2-15 

2.9 Instrumentation 2-17 

3.0 TEST RESULTS 3-1 

3.1 Introduction 3-1 

3.2 Steady-State Test Results 3-3 

3.2.1 Seal Set No. 1 3-3 

3.2.2 Seal Set No. 2 3-6 

3.2.3 Seal Set No. 3 3-10 

3.2.4 Seal Set No. 4 3-11 

3.3 Dynamic Behavior During Seal Testing 3-17 

3.3.1 Seal Set No. 2 3-18 

3.3.2 Seal Set No. 3 3-19 

3.3.3 Seal Set No. 4 3-20 

3.4 Acceleration Testing 3-22 

3.4.1 Seal Set No. 2 3-22 

3.4.2 Seal Set No. 3 3-23 



'>^)fJKDfNG PAGE BLANK NOT FJLMED 



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pMI I \/ INTFwnONAllT tHAM 



-V- 



TABLE OF CONTENTS (CONTINUED ) 

SECTION PAGE 

3.5 Post-Test Hardware Inspections and Failure Analysis .... 3-23 

3.5.1 Seal Set No. 1 3-24 

3.5.2 Seal Jet No. 2 3-25 

3.5.3 Seal Set No. 3 3-26 

3.5.4 Seal Set No. 4 3-27 

3.6 Discussion Of The Results . 3-27 

3.6.1 Steady-State Operation . 3-28 

3.6.2 Start-Up Performance 3-29 

3.6.3 Seal Life 3-30 

3.6.4 Leakage Races 3-32 

3.6.5 Parametric Effects 3-33 

3.6.6 Dynamic Performance 3-41 

3.6.7 Material Considerations 3-42 

4.0 ANALYSIS AND DESIGN OF RAYLEIGH-STEP, HuLIUM BUFFER SEALS .... 4-1 

4.1 Operating Conditions 4-1 

4.2 Design Considerations 4-1 

4.3 Analysis and Design of the 50-mm Floating-Ring Helium 

Purge Seal 4-2 

4.3.1 Rayleigh Step Optimization Studies 4-2 

4.3.2 General Configuration and Design 4-4 

4.3.3 Fluid-Film Performance 4-5 

4.3.4 Thermal Analysis 4-7 

4.3.5 Seal Ring Dynamic Response 4-9 

4.3.6 Summary of Results and Conclusions of 

Analytical Studies 4-12 

4.4 Analysis and Design of the 20-mm Floating-Ring Helium 

Purge Seal 4-13 

4.4.1 General Configuration and Operating Conditions . . . 4-13 

4.4.2 Fluid-Film Performance „ 4-13 

4.4.3 Dynamic Response < 4-15 

5.0 ANALYSIS AND DESIGN OF SPIRAL-GROOVE LOX SEALS 5-1 

5.1 General Discussion 5-1 

5.2 Analytical Approach 5-1 

5.3 Conf icuration for the 50-mm Seal 5-2 

5.4 Calculated Fluid-Film Performance of the 50-mm 

Spiral-Groove Seal 5-4 

5.5 Dynamic Analysis of the 50-mm Spiral-Groove Seal 5-6 

5.6 Elastic and Thermal Distortions of the 50-mm 

Spiral-Groove Seal 5-8 

5.7 Design of the 20-iiun Spiral-Groove Seal 5-9 

6.0 TEST RIG „ 6-1 

6.1 General Configuration 6-1 

6.2 Fluid Systems 6-2 



-vi- 



TABLE OF CONTENTS (CONTINUED) 



SECTION 



PAGE 



6.3 

6.4 



6.5 



6.6 



Test Seal Arrangement ^ . 

Turbine Design and Performance 

6.4.1 Acceleration , . . . 

6.4.2 Stress 

6.4.3 Design 

Bearing Design and Performance 

6.5.1 Journal Bearings 

6.5.2 Thrust Bearings 

Rotor Dynamics 

6.6.1 Rotor Model for the 50-mm LOX Seal Shaft . . . 

6.6.2 Undamped Critical Speeds for the 50-mm LOX 

Seal Rotor 

6.6.3 Synchronous Unbalance Response of the 50-mm LOX 

Seal Rotor 

6.6.4 Stability Analysis for the 50-tnm LOX Seal Rotor 



6.7 Thermal Analysis 



6-5 
6-5 
6-6 
6-7 
6-7 
6-7 
6-7 
6-8 
6-9 
6-12 

6-12 

6-13 
6-14 
6-15 



7.0 TEST FACILITY 

7.1 Fluid Supply Systems 

7.1.1 Helium Seal Supply System 

7.1.2 LN2 Bearing Supply System 

7.1.3 GN2 Turbine Supply System 

7.1.4 Helium Supply to Labyrinth Seal 

7.1.5 GN2 Purge 

7.2 Controls 

7.3 Instrumentation 

7.3.1 Film Thickness and Shaft Displacement . . . . 

7.3.2 Seal Film Thickness Probes 

7.3.3 Shaft Displacement Probes 

7.3.4 Shaft Speed 

7.3.5 Vibration 

7.3.6 Pressures 

7.3.7 'lows 

7.3.8 Temperatures 

7.3.9 Data Acquisition Equipment 

8.0 TEST PLAN FOR THE RAYLEIGH-STEP, HELIUM BUFFER SEAL . . . 

8.1 Test Description 

8.1.1 Sceady-State Tests 

8.1.2 Acceleration Tests 

8.2 Test Schedules 

8.3 Test Procedures 

8.3.1 Preparation of Test Rig at MTI 

8.3.2 Preparation of Test Facility and Test Rig at 

Wyle Laboratories 

8.3.3 Test Facility Operating Procedures 

8.3.4 Pre- and Post-Test Inspection and 

Assembly Procedures . 

8.3.5 Instrument Calibration , 

8.3.6 Data Reduction 



7-1 

7-1 

7-1 

7-2 

7-4 

7-5 

7-5 

7-6 

7-9 

7-9 

7-10 

7-12 

7-12 

7-12 

7-13 

7-13 

7-13 

7-13 

8-1 

8-1 

8-1 

8-4 

8-5 

8-5 

8-5 

8-6 
8-6 

8-11 
8-11 
8-12 



9.0 



REFERENCES 



9-1 



-vii- 



LIST OF FIGURES 

MUMBER PAGE 

1-1 LOX Turbopump Cross Section 1-5 

2-1 Floating-Ring Seal Schematic 2-18 

2-2 Developed View of 50-mm Rayleigh-Step Pad 2-19 

2-3 50-nm Rayleigh-Step Floating Rings 2-20 

2-4 Operating Map for 50-min Helium Buffer Seal - 

Seal Set No. 2 2-21 

2-5 Outboard Seal Flow, 103 kPa Drain Pressure - 

Seal Set No. 2 2-22 

2-6 Zero Speed versus Pressure Drop - Seal Set No. A 2-23 

2-7 Flow versus Pressure Drop - Seal Set No. A 2-23 

2-8 Seal Leakage Envelope Data at Varying Speeds 2-2A 

2-9 Seal Leakage "Envelope Data at Constant Speed 2-25 

2-10 Operating Range Map tor 20-mm Helium Buffer Seal 2-26 

2-11 Pressure-Balanced, Outward Pumping, Spiral-Groove Concept . 2-27 

2-12 Seal Face Pressure Profile 2-28 

2-13 Groove G^aometry 2-29 

2-14 50-rani Pressure-Balanced LOX Seal 2-30 

2-15 50-mm Spiral-Groove Seal, Mating Ring, and Face Seal . . . 2-31 

2-16 Operating Range Map for 50-mm Spiral-Groove Seal 2-32 

2-17 20-nwi Pressure-Balanced LOX Seal 2-33 

2-18 Operating Range Map for 20-mm Spiral-Groove Seal 2-34 

2-19 Test Rig Cross Section ..." 2-35 

2-20 Test Rig Components 2-36 

2-21 Helium Seal Rotor 2-37 

2-22 Modified Test Rig Configuration - Seal Set No. 4 ' 2-38 

2-23 Details of Embedded Probe Installation 2-39 

2-24 Instrumented Seal Ring with Embedded Capacitance Probes . . 2-40 

3-1 Pressure History - Seal Set No. 1 3-48 

3-2 Speed History - Seal Set No. 1 3-48 

3-3 Operating Map for Seal Set No. 1 3-49 

3-4 Failure of No. 1 Outboard Seal Ring 3-50 

3-5 Zero Speed Flow versus Pressure Drop - Seal Set No. 1 . . . 3-51 
3-6 Zero Speed Seal Temperatures versus Pressure Drop - 

Seal Set No. 1 3-51 

3-7 Flow versus Pressure Drop - Seal Set No. 1 3-52 

3-8 Inboard Seal Temperature versus Pressure Drop - 

Seal Set No. 1 3-53 

3-9 Outboard Seal Temperature versus Pressure Drop - 

Seal Set No. 1 3-53 

3-10 Inboard and Outboard Seal Temperature versus Presure 

Drop - Seal Set No. 1 3-54 

3-11 First Day Pressure History - Seal Set No. 2 3-55 

3-12 First Day Speed History - Seal Set No. 2 3-55 

3-13 Second Day Pressure History - Seal Set No. 2 3-56 

3-14 Second Day Speed History - Seal Set No. 2 3-56 

3-15 Third Day Pressure History - Seal Set No. 2 3-57 

3-16 Third Day Speed History - Seal Set No. 2 .... 3-57 

3-17 Operating Map for Seal Set No. 2 3-58 

3-18 Inboard Seal Flow versus Pressure Drop, Zero Speed - 

Seal Set No. 2 3-59 

UM,JjUjUJMimOHm.M nm -ix- gj^^jt^j^u^G p^ge P.ANK not tOMID 



LIST OF FIGURES (CONTINUED) 

3-19 Outboard Seal Plow versus Pressure Drop, Zero Speed - 

Seal Set No. 2 3-59 

3-20 Zero Speed Seal Temperatures versus Pressure Drop - 

Seal Set No. 2 3-60 

3-21 Inboard Seal Flow versus Pressure Drop - Seal Set No. 2 . . 3-61 
3-22 Inboard Seal Temperature versus Pressure Drop - 

Seal Set No. 2 3-61 

3-23 Outboard Seal Flow, 517 kPa Drain Pressure - 

Seal Set No. 2 3-62 

3-24 Outboard Seal Flow, 310 kPa Drain Pressure - 

Seal Set No. 2 3-63 

3-25 Outboard Seal Temperature, 517 kPa Drain Pressure - 

Seal Set No. 2 3-64 

3-26 Outboard Seal Temperature, 310 kPa Drain Pressure - 

Seal Set No. 2 3-65 

3-27 Outboard Seal Temperature, 103 kPa Drain Pressure - 

Seal Set No. 2 3-66 

3-28 Outboard Seal Temperature and Flow versus Pressure 

Drop - Seal Set No. 2 , 3-67 

3-29 Outboard Seal Temperature and Flow versus Supply- 
Pressure - Seal Set No. 2 3-67 

3-30 First Day Seal Film Thickness Histor^ il Set No. 2 . . 3-68 
3-31 Second Day Film Thickness History - Se .oit No. 2 .... 3-69 

3-32 Third Day Film Thixkness History - Seal Set No. 2 3-70 

3-33 Pressure History - Seal Set No. 3 3-71 

3-34 Speed History - Seal Set No. 3 3-71 

3-35. Operating Map - Seal Set No. 3 • 3-72 

3-36 Failure of No. 3 Inboard Seal Ring 3-73 

3-37 Inboard Seal Flow versus Pressure Drop - Seal Set No. 3 . . 3-74 
3-38 Inboard Seal Temperature versus Pressure Drop - 

Seal Set No. 3 3-74 

3-39 Outboard Seal Flow, 517 kPa Drain Pressure - 

Seal Set No. 3 3-75 

3-40 Outboard Seal Flow, 310 kPa Drain Pressure - 

Seal Set No. 3 3-76 

3-41 Outboard Seal Flow, 103 kPa Drain Pressure - 

Seal Set No. 3 3-77 

3-42 Outboard Seal Temperature, 517 kPa Drain Pressure - 

Seal Set No. 3 3-78 

3-43 Outboard Seal Temperature, 310 kPa Drain Pressure - 

Seal Set No. 3 3-79 

3-44 Outboard Seal Temperature, 103 kPa Drain Pressure - 

Seal Set No. 3 3-80 

3-45 Operating Map of Seal Set No. 4 3-81 

3-46 Pressure History - Seal Set No. 4 3-82 

3-47 Speed History - Seal Set No. 4 3-82 

3-48 Failure of No. 4 Inboard Seal Ring 3-83 

3-49 Flow History - Seal Set No. 4 3-84 



-X- 



LIST OF FjGURES^CaM -INUED) 

3=50 Seal Temperature History - Seal Set No. 4 3-85 

3=51 Outboard Seal Film Thickness History - Seal Set No. 4 . . . 3-86 

3-52 Outboard Seal Eccentricity History - Seal Set No. 4 . . . , 3-86 

3-53 Outboard Seal Film Thickni- a versus Shaft Speed - 

Seal Set No. 4 3-87 

3-54 Seal and Runner Motion - Seal Set No. 2 at 3665 rad/s . . . 3-88 

3-55 Seal and Runner Motion - Seal Set No. 2 at 4188 rad/s . . . 3-88 

3-56 Seal and Runner Motion - Seal Set No. 2 at 4712 rad/s . . . 3-88 

3-57 Seal and Runner Motion - Seal Set No. 3 at 3665 rad/s . . . 3-89 

3"-58 Seal and Runner Motion - Seal Set No. 3 at 3665 rad/s . . . 3-89 

3-59 Seal and Runner Motion - Seal Set No. 3 at 3665 rad/s . . . 3-90 

3-60 Seal and Runner Motion - Seal Set No. 3 at 3665 rad/s . . . 3-90 

3-61 Seal and Runner Motion - Seal Set No. 3 at 4188 rad/a . . . 3-91 

3-62 Seal and Runner Motion - Seal Set No. 3 at 4712 rad/s . . . 3-91 

3-63 Seal and Runner Motion - Seal Set No. 4 at 3665 rad/s , . . 3-92 

3-64 Seal and Runner Motion - Seal Set No. 4 at 4188 rad/s . . . 3-92 

3-65 Seal and Runner Motion - Seal Set No. 4 at 4712 rad/s . . . 3-92 

3-66 Seal and Runner Motion - Seal Set No. 4 at 5235 rad/s . . . 3-93 

3-67 Seal and Runner Motion - Seal Set No. 4 at 5759 rad/s . . . 3-93 

3-68 Embedded Probe Oscilloscope Display Format 3-94 

3-69 Seal Probe Time Trace - Seal Set No. 4 at 5759 rad/s . . . 3-95 

3-70 Speed-Time Curve for Typical Acceleration Run 3-96 

3-71 Damaged Seal; Outboard Ring - Seal Set No. 1 3-97 

3-72 Undamaged Seal; Inboard Ring - Seal Set No. 1 3-9/ 

3-73 Damaged Runner; Inboard Left, Outboard Right - 

Seal Set No. 1 3-98 

3-74 Magnified View (Xll.4) of Damaged Runner - Seal Set No . 1 . 3-98 

3-75 Undamaged Seal; Outboard Ring - Seal Set No. 2 3-99 

3-76 Damaged Seal; Outboard Ring - Seal Set No. 3 3-100 

''-77 Undamaged Seal; Inboard Ring - Seal Set No. 3 3-100 

3-78 Damaged No. 2 Runner; Inboard Left, Outboard Right - 

Seal Set No. 3 3-101 

3-79 Magnified View (X11.5) of Damaged No. 2 Runner 3-101 

3-80 Damaged Seal; Inboard Ring - Seal Set No. 4 3-102 

3-81 Damaged No. 3 Runner; Inboard Left, Outboard Right - 

Seal Set No. 4 3-102 

3-82 Parametric Variations as a Function of Pressure Drop 

and Supply Pressure 3-103 

4-1 Dimensionless Direct Stiffness versus Step Depth 4-22 

4-2 Axial Step-Length Optimization 4-23 

4-3 Optimization of Circumferential Pocket Extent 4-24 

4-4 50-mm Rayleigh-Step Pad 4-25 

4-5 50-nnm Runner Seal Installation 4-26 

4-6 50-ntim Helium Seal Runner Centrifugal Distortions 4-27 

4-7 50-mm Helium Purge Bushing Seal Assembly 4-29 

4-8 Bushing Seal, CCW 4-31 

4-9 Concentric Clearance versus Speed Including Centrifugal 

Growth 4-33 

4-10 50-mm Fluid-Film Radial Force and Seal Friction Force . . . 4-34 



-xi- 



LI STJ)£^ FIGURES (CONTPJl'KP) 

4-11 50-mm Vijjcoua Power Loss as a Function of Speed 

and Pressure, e ^ q.S 4-35 

4-12 50-mm Seal Leakage verauo Spend and Pressure 4-36 

4-13 50-mm Fluid Temperature Rise 4-3/ 

4-14 50-mm Thermal Analysis Model 4-38 

4-15 50-mm Fluid Film Forces versus Eccentric' y Ratio; 

1379 kPa Helium Pressure; 7330 rad/s 4-39 

4-16 50-mm Ring Orbital Response; 1379 kPa, 7330 rad/s, 

7.62 (im Shaft Runout 4-40 

4-17 50-mm Ring Orbital Response, 345 kPa Absolute (50 psia), 

2094 rad/s (20,000 r/min), 0.0102 mm (0.0004 in.) 

Shaft Runout 4-41 

4-18 50-mm Ring Orbital Response; 689 kPa Absolute (100 psia); 

2094 rad/s (20,000 r/min); 0.0102 mm (0.0004 in.) 

Shaft Runout 4-42 

4-19 50-mm Ring Orbital Response; 1379 kPa Absolute (200 puia); 

3142 rad/s (30,000 r/min); 0.0254 mm (0.0004 in.) 

Shaft Runout 4-43 

4-20 50-mm Composite Mv-l jl Ring, Orbital Response; 1379 kPa 

Absolute (200 psia), 7330 rad/s (70,000 r/min); 

0.00635 mm (0.00025 in.) Shaft Runout 4-44 

4-21 50-mm Ring Resoonse versus Shaft Orbit 4-45 

4-22 50-rnn Composite and All Carbon Ring Orbit Summary 

versus Shaft Runout . 4-46 

4-23 20-mm Helium Purge Seal 4-47 

4-24 20-mm Helium Seal; Force versus Eccentricity at 

Different Pressures and Speeds ..... 4-49 

4-25 20-mm Helium Seal; Pressure versus Eccentricity to 

Balance Frictional Force 4-50 

4-26 20-mm Helium Seal; Effect of Clf;arance on Fluii Film 

Force; 10,472 rad/s (100,000 r/min); 1379 kPa 

Absolute (200 psia) 4-51 

4-27 20-mm Helium Seal; Force versus Eccentricity at Larger 

Clearance (CO = 0.0254 mm) 4-52 

4-28 20-mm Helium Buffer Seal at 0.5 Eccentricity Ratio; 

Leakage versus Helium Pressure 4-53 

4-29 20-mm Helium Seal; Leakage versus Eccentricity at 

1379 kPa (200 psia) 4 54 

4-30 20-mm Helium Seal, Power Los3 versus Speed 4-55 

4- j1 20-mm Helium Buffer Seal at 0.5 Eccentricity Ratio; 

Temperature Rise versus Speed 4-56 

4-32 20-mm Helium Seal, Fluid Film Force versus Eccentricity 

at 1379 kPa Absolute (200 psi), 10,472 rad/s 

(100,000 r/min) 4-57 

4-33 20-mm Helium Seal; Transient Response Summary; 

0.0025 mm Shaft Runout 4-58 

4-34 20-mm Helium Seal; Transient Response Summary versus 

Buffer Pressure; 0.0025 mm Shaft Runout 4-59 



-xii- 



l^ 



' ST 0['\ FI GURES_(CONTI NUED ) 

5-1 50-nm LOX Spiral-Groove Seal Asaembly 5-11 

5-2 50-mm Spiral-Groove Face Seal 5-13 

5-3 SO-mm Spiral-Groove Seal Mating Ring 5-15 

5-4 50-inra Spiral-Groove Secondary Seal 5-17 

5-5 Split Spiral-Groove Secondary S«.^al 5-19 

5-6 50-mm Spiral-Groove Seal; Film Th'ckneaa versus 

Speed and Pressure 5-20 

5-7 50-niin Spiral-Groove Seal; Axial Stiffness versus 

Speed and Pressure 5-21 

5-8 50-mm Spiral-Groove Seal; Spiral Groove Circulating Flow 

versus Spepl and Pressure 3-22 

5-9 50-mm Spiral-Groove Seal; Leakage Flow versus Speed 

and Pressure 5-23 

5-10 50-mm Spiral-Groove Seal; Power Loss versus Speed 

and Pressure 5-24 

5-11 50-mm Spiral-Groove Seal; Spiral-Groove Temperature Rise 

versus Speed and Pressure 5-25 

5-12 50-mm Spiral-Groove Seal; Temperature Rise versus / 

Speed and Pressure 5-26 | 

5-13 50-mm Spiral-Groove Seal; Angular Amplitude Ratio 

versus Speed and Pressure 5-27 

5-14 50-mm Spiral-Groove Seal; Thermoelastic Distortion .... 5-28 

5-15 2Q-!tHT! Spiral-Groove Seal; Film Thickness versus Speed . . . 5-29 

5-16 20-mm Spiral-Groove Seal; Axial Stiffness versus Speed . . 5-30 

5-17 20-mm Spiral-Groove Seal; Leakage Flow versus Speed . . , . .5-31 

5-10 20-mm Spiral-Groove Seal; Groove Flow versus Speed .... 5-32 

5-19 20-mm Spiral-Groove Seal; Power Loss versus Speed 5-33 ;| 

5-20 20-mm Spiral-Groove Seal; Seal Temperature Rise ', 

versus Speed 5-34 > 

5-21 20-mm Spiral-Groove Seal; Seal Teirperature Rise ^ 

versus Speed 5-35 * 

5-22 20-mm Spiral-Groove Seal; Axial Natural Frequency 

versus Speed 5-36 M 

5-23 20-mm Spiral-Groove Seal; Angular Nati ral Frequency * 

versus Speed 5-37 

5-24 20-mm Spiral-Groove Seal; Axial Amplitude Ratio 

versus Speed 5-38 

5-25 20-mra Spiral-Groove Seal; Angular Amplitude Ratio 

versus Speed 5-39 

"-26 Spiral-Groove Seal - Elastic Distortions 5-40 

6-1 Cross Section of Seal Test Rig; Housing Sections 

Identified 6-16 

6-2 Test Rig Housings 6-17 

6-3 Test Rig Nozzle Box 6-18 

6-4 Test Rig Fluid Systems 6-19 

6-5 Supply and Drainage - Turbine-Side Bearing Housing 6-20 

6-6 Supply and Drainage - Seal-Side bearing Housing 6-21 

6-7 Supply and Drainage - Labyrinth Seal and Shim Plate 

Housings 6-22 



-xiii- 



LI ST OJ^ F I CURES ( CONTI NUED ) 

5-8 Teat Rig Power Losses b-2^ 

6 9 Turbine Design Schematic , . . b-ift 

6 Assembled Test Rig Shaft <i-2'j 

6-11 Turbine Output Power versus Inlet Pressure 6-2^ 

b-12 Turbine Efficiency Ratio versus Speed b-27 

6-13 Turbine Wheel 6-2') 

6-14 Test Rig Journal Bearing Recess Geometry 6=31 

6-15 Test Rig Thrust Bearing Recess Geometry 6-32 

6-16 Test Rig Thrust Loading 6-33 

6-17 Thrust Bearing Load Capacity versus Displacement 6-34 

6-18 Thrust Bearing Flow versus Displacement 6-35 

6-19 Thrust Bearing Power Loss - Single Side 6-36 

6-20 Thrust Bearing Recess Pressure 6-37 

6-21 Thrust Bearing Temperature Rise versus Film Thieknass . . . 6-38 

6-22 Rotordynamies Model of Test Rig 6-39 

6-23 Undamped Critical Speed Map - 50-mm LOX Seal Shaft .... 6-40 
6-24 Half Amplitude Response, Slationa 8, 10, 11; 

In-Phase Unbalance, C = 0.0191 mm 6-41 

6-25 50-mm Seal Half-Amplitude Response, Stations 9, 10, 15; 

In-Phase Unbalance, G = 0,0191 mm 6-42 

6-26 50-mra Seal Half-Amplitude Response, Stations 1, 16, 20; 

In-Phase Unbalance, C = 0.0191 mm 6-43 

6-27 50-ffiffl Seal Half-Amplicude Response, Stations 8, 10, 11; 

Out-of-Phase Unbalance, C = 0.0191 mm , j-44 

; • 3 50-mm Seal Half-Aiwplitude Response, Stations 9, 10, 15; 

Out-of-Phase Unbalance; C = 0.0191 mm 6-45 

6-29 50-mni Seal Half-Amplitude Response, Stations 1, 16, 2Gj 

Out-of-Phase Unbalance, C = 0.0191 mm 6-46 

6-30 LOX Seal Test Rig 6-47 

6-31 Test Rig Drain Pressure versus Saturation Temperature . . . 6-48 

7-1 Overall View of Helium Seal Test Facility 7-15 

7-2 Close-Up View of LN2 Manifold and Bearing Supply Valves . . 7-16 

7-3 Close-Up View of Test ■'.ig Drain Valves 7-17 

7-4 Close-Up View of Helium Seal Test Rig 7-18 

7-5 Overhead View of Helium jeal Test Rig 7-19 

7-6 Overall View of Control Room 7-20 

7-7 Close-Up View of System Control Panel 7-21 

7-8 Fluid Supply System Schematic 7-23 

7-9 Test Rig Instrumentation , 7-25 

7-10 Capacitance Probe Construction 7-26 

7-11 50-mra Helium Seal Probe Configuration .... 7-27 

7-12 Capacitance Probes, LOX Seal, and Shaft Probes 

Configurations , 7-28 

7-13 Simplified Schematic of Data Acquisition Equipment .... 7-29 

8-1 Typical Seal Operating Sequence 8-14 

8-2 Typical Data Logger Scan from Seal Test No. 4 8-15 

8-J Typical Computer-Ge.ierated Plot . . o 8-16 



-XxV- 



LIST OF TABLES 



NUMBER 



PAGE 



1-1 REQUIREMENTS AND OPERATING CONDITIONS 1-C 

2-1 SUMMARY OF TEST RESULTS 2-41 

2-2 DIMENSIONS OF PRESSURE-BALANCED, 50-MM, OUTWARD PUMPING, 

SPIRAL-GROOVE SEAL 2-42 

2-3 STEADY-STATE PERFORMANCE OF 50-MM SPIPAL-GROOVE SEAL ... 2-43 
2-4 DIMENSIONS OF PRESSURE-BALANCED, 20-MM, OUTWARD PUMPING, 

SPIRAL-GROOVE SEAL 2-44 

2-5 STEADY-STATE PERFORMANCE OF 20-MM SPIRAL-GROOVE SEAL . . . 2-45 

3-1 SEAL DIMENSION SUMMARY 3-44 

3-2 ACCELERATION RUN PERFORMANCE DATA - SEAL SET NO. 2 .... 3-45 

3-3 ACCELERATION RUN PERFORMANCE DATA - SEAL SET NO. 3 .... 3-46 

3-4 SEAL INSPECTION SUMMARY 3-47 

4-1 PROPERTIES OF HELIUM 4-17 

4-2 REQUIRED ECCENTRICITY AND F LM THICKNESS TO OVERCOME 

FRICTIONAL RESISTANCE 4-18 

4-3 MAXIMUM OPERATING SPEED FOR tX '^ 22° C 4-18 

4-4 RESULTS OF 50-MM THERMAL ANALYSIS 4-19 

4-5 RECOMMENDED DIMENSIONS ACCOUNTING FOR CENTRIFUGAL 

GROWTH AND THERMAL CONTRACTIONS 4-19 

4-6 50-MM SEAL TFJ^NSIENT ANALYSIS: SUMMARY OF RESULTS 4-20 

4-7 20-MM SEAL TRANSIENT ANALYSIS: SUMMARY OF RESULTS 4-21 

5-1 50-MM PRESSURE-BALANCED SEAL 5-41 

5-2 20-MM PRESSURE-BALANCED SEAL 5-42 

6-1 SUMMARY OF CONCENTRIC JOURNAL BEARING PERFORMANCE 6-49 

6-2 SUMMARY OF THRUST BEARING PERFORMANCE (7329 RAD/S; 

ORIFICE (do = 2.08 MM) 6-50 

6-3 SUMMARY OF THRUST BEARING PERFORMANCE (10,470 RAD/S; 

ORIFICE (do = 1.93 MM) 6-51 

6-4 50-MM LOX SEAL ROTOR PARAMETERS 6-52 

6-5 50-MM LOX SEAL STABILITY ANALYSIS 6-53 

7-1 INSTRUi'- .'TATION MATRIX 7-30 

7-2 MTI ACCUMEASURe"^^ •'YSTEM 1000 REQUIREMENTS AND 

SPECIFICATIONS , 7-31 

8-1 SEAL SET NO. 1 STEADY-STATE TEST SCHEDULE 8-17 

8-2 SEAL SET NO. 2 STEADY-STATE TEST SCHEDULE 8-19 

8-3 SEAL SET NO. 3 STEADY-STATE TEST SCHEDULE 8-22 

8-4 SEAL SET NO. 4 STEADY-STATE TEST SCHEDULE 8-24 

8-5 SEAL SET NO. 2 ACCELERATION TEST SCHEDULE 8-26 



-XV- 



1.0 ^.liJTRQQUCTION 

Liquid oxygen (LOX) turbopumpa are high-speed, high-presgure, and high-power 
density machines that require effective sealing between the high-pressure LOX 
being pumped at the impeller end of the machine and the gaseous steam at the 
turbine end of the unit. 

Under NASA Contract NAS3-23260, Mechanical Technology Incorporated (MTI) 
conducted a technology program for <".ealg to be used in advanced LOX turbo- 
pumps. Two types of seals were examined; a primary spiral-groove LOX ueal and 
a floating-ring helium buffer seal with Rayleigh step, hydrodynamic lift pads. 
Two shaft sizes of 20 mm and 50 mm were considp ^ for each seal type. The 
function of the spiral-groove LOX seal is to limit leakage of high-pressure 
LOX generated by the pump impeller. It seals LOX fluid in a seal chamber in 
which the pressure is typically maintained at 3 to 5 MPa (435 to 725 psig). 
The function of the floating-ring helium buffer seal is to separate the gase- 
ous steam that drives the turbine from the LOX end of the machine. Helium is 
used as a buffer fluid at a maximum pressure of 1.38 MPa (200 psia). 

Figure 1-1* is a cross section of a LOX turbopump that incorporates both types 
of seals. Other configurations that utilize a double suction impeller obviate 
the need for the LOX end seal, but still require the helium buffer seal. 

In Figure 1-1, the flow path through the spiral-groove face seal is as 
follows. From the discharge of the impeller, the flov; that does not enter the 
volute leaks to the backside of the impeller and into the bearing compartment 
through two orifices formed by the clearance space between the impeller and 
the housing. These orifice openings adjust as a function of the thrust load- 
ing and direction and form an integral component of the thrust balancing 
system. After flowing through the bearings (for cooling purposes), the flow 
passes through a labyrinth seal to break down the pressure from approximately 
29.6 MPa to 3.45 MPa maintained In the seal chamber. Upon entering the seal 



*For ease of readership, illustrations throughout this report are included at 
the end of each section. In this section, they begin on page 1-5. 



l-l 



chamber, the flow takes two paths. One path i i the leakage through the face 
type seal; the other is a recirculation line to the pump inlet. These sepa- 
rate flow paths are indicated on Figure 1-1. It is important that there is a 
constant injection flow circulating through the seal chamber to carry away 
heat developed by the "^ace seal. 

At the turbine end of the machine, helium buffer fluid is introduced between 
the opposed pair of rings of the floating-ring seals at a maximum pressure of 
approximately 1379 kPa (200 psia). The leakage from the face seal and LOX 
side of the floating-ring seal discharge from the machine through a common 
drain. 

The seal development program was not directed toward a specific machine, but 
was intended to cover a broad range of sizes and conditions. Two seal sizes 
were investigated, one for a 20-mm diameter shaft and the other for a 50-nim 
diameter shaft. Table 1-1 identifies targeted requirements and operating 
conditions . 

Although leakage recfuirements were not specified, they wei = to be as small as 
practicable because leakage reduction translates directly inro increased 
vehicle payload. The 183 m/s (600 ft/s) interface speed was a target require- 
ment that could not practically be met with the 20-mm shaft diameter. It 
necessitates operating shaft speeds of approximately 18,326 rad/s 
(175,000 r/min). The maximum design speed of the rig was 10,472 rad/s 
(100,000 r/min). 

The program that was accomplished included the following: 

1. Analysis, design, manufacture, and test of 50-mm Rayleigh-step helium 
buffer seals . 

2. Analysis and design of 20-mm Rayleigh-step helium buffer seals. 

3>. Analysis, design, and manufacture of 50-mm spiral-groove LOX seals. 
4. Analysis and design of 20-mm spiral-groove LOX seals. 



1-2 



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5. Analysis, design, and manufacture of a seal test rig designed fr ■ maximum 
speeds of 7330 rad/s (70,000 r/min) for 50-mm seal testing and 
10,472 rad/s (100,000 r/min) for 20-mm seal testing. 

6. Design and installation of a sf:al test facility at Wyle Laboratories, 
Norco, California. 

7. Completion of a test program for the 50-mm, helium buffered Rayleigh-step 
seal . 

8. The program produced extensive documentation. In addition to this final 
report and separate Executive Summary, References [l]*, [2], and [3] are 
interim reports concerned with the analysis and design of the LOX spiral- 
groove seals, Rayleigh-step helium buffer seals, and the test rig, respec- 
tively. Reference [4] describes the test plan including the design of the 
test facility. 



Several sets of spiral-groove face seals and Rayleigh-step floating ring seals 
were manufactured in the 50-mm size, by Stein Seal Company of Philadelphia, 
a Pennsylvania. Stein completed manufacturing drawings and contributed to the 



final designs. 

In addition to the tasks described above, MTI supplied to NASA/LeRC for inter- 
nal use a series of computer codes for analyzing the spiral-groc /e and 
Rayleigh-step seals. Brief descriptions of these codes follow. 



Spiral-Groove Seal Codes 

SPIRALTI - is used for establishing the geometry and fluid-film performance 
of spiral-groove seals. Turbulence and ine-tia are included and the geom- 
etry can be optimized on the basis of stiffness. 



^Numbers in brackets denote references that can be found in Section 9.0. 



1-3 



DSEALBI2 - establishes complete seal performance of a given geometry includ- 
ing film thickness, flow, power loss, temperature rise, elastic thermal 
distortions and stresses, and dynamic response characteristics. 

* Ploating-Ring Seal Codes 

RASTEPCO - is used for optimizing the geometry of the shrouded Rayleigh-step 
compressible fluid seal configuration and producing steady state perform- 
ance. The code alsf produces fluid-film forces as a function of eccentrici- 
ty vatio for use in the dynamics code described below. 

RINGDY - determines dynamic response of a fluid-film (compressible fluid) 
floating-ring seal as a function of given shaft excursions. Coulomb fric- 
tion occurring along the side walls is taken into account. The program 
produces "rbital response of the ring and establishes whether the ring is 
properly following the shaft. 

FLOWCAL - was written to provide more accurate leakage predictions for the 
floating-ring seal. The RASTEPCO code computes flow on the basis of viscous 
laminar theory. FLOWCAL computes gas flow through an annular clearance of 
finite length exposed to an axial pressure gradient, including combined 
inertia, viscous effects, and seal ring eccentricity. The thermodynamic 
behavior of the fluid can be preselected to be isothermal or adiabatic. 

This report has been organized so that significant results, conclusions, and 
recommendations are presented up front. For those involved in the details of 
seal technology including analysis, design, testing, instrumentation, etc., 
subsequent sections of the report would be informative. Also, an Executive 
Summary (MTI 85TR21) has been generated that concisely presents the more 
significant results of the program effort. 



1-4 



Balance Piston 
Orifices 



Pressure Breandown 
Labyrinth 



Recirculating Line 
to Pump Inlet 



Pump 
Impeller 



Turbine 




Rotating 
Mating Ring 



Shaft 



Floatinq-Rinq 
Buffered Seals 



Fig. 1-1 LOX Turbupump Cross Section 



tl3MS 



I >Mi>iai*iiifliiiriii aim nn -iniia— f inrfi 



TABLE 1-1 
REQUIREMENTS AND OPERATIN G CONDITIONS 

Spiral Groove Seals 

Shaft Diameter (mm): SO/20 

Maximum Pressure (MPa): S. 17/5. 17 (750 psia) 

Fluid: un 

Interface Speed: 183 m/s (600 ft/s^) 

Start-Up Acceleration: 152.4 m/s^ (500 £t/s^) 

Minimum Operating Life (h): 10 

Minimum Number of Starts: 300 

Helium Buffer Seals 

Shaft Diameter (mm): 50/20 

Maximum Pressure (MPa): 1.38/1.38 (200 psia) 

Fluid: Helium Gas 

Interface Speed: 183 m/s (600 ft/s) 

Start-Up Acceleration: 152.4 m/s^ (500 ft/s^) 

Minimum Operating Life (h): 10 

Minimum Number of Starts: 300 



LI 



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1-6 



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2.0 SUMMARY OF SIG NIFICANT RESULTS AND CONCLUSIONS 

In this aection, a broad summary of the results of the program are presented 
with relevant descriptive material, conclusions, and recommendations. Both 
types of seals are discussed as well as the test rig and instrumentation. 

2.1 Configu r ation and Principle of Operation of a 50-mm Helium Buffer Seal 

Figure 2-1* shows a schematic representation of the helium buffered, float- 
ing-ring seal. The seal consists of two rings that are mounted back-to-back. 
The helium buffer fluid enters between the rings and forces the rings up 
against the stationary housing. The buffer fluid leaks in the clearance annu- 
lus between the shaft and the seal and prevents ingress of exterior fluid on 
either side of the floating-ring assembly. 

The rings are held in equilibrium by a number of forces as shown in 
Figure 2-1. Fc is a pressure force from the inlet buffer fluid that forces 
the rings up against the housings. This pressure force is balanced pafrt way 
on the housing sides of the rings indicated by Fg. Fpj represents a hydrody- 
namic force that is generated by rotation between the shaft and the ring. The 
net hydrodynamic force is zero when the shaft and rings are in the concentric 
position. However, when Che ring becomes eccentric with respect to the shaft, 
a hydrodynamic force is built up that opposes the eccentricity. There is also 
a normal force, F^, acting on the ring at the contact area bet .'een the ring and 
the housing. Generally, F^j is maintained as small as possible to minimize 
frictional resistance forces. To minimize F^, the balance force, Fg, should 
be as large as possible. Therefore, the contact area is small and is located 
as close to the shaft as is practicable, but with sufficient housing clearance 
to permit the necessary shaft excursions. In addition to the equilibrium 
forces mentioned above, there is a friction force, Ff, between the seal ring 
and housing. 



*Figures are presented consecutively, beginning on page 2-18. 



2-1 



Figure 2-2 shows the hydrodynamic geometry that is incorporated into the bore 
of the seal rings. A portion of the length of the bore ig segregated into 
sectors and these sectors arc separated from one another by axial grooves. A 
circumferential groove that goes completely around the bore is incorporated 
before the seal dam region. At the interior of the sectors, Rayleigh-step 
pockets are machined. Hydrodynamic pressures are generated by viscous pumping 
of the fluid over the Rayleigh step. Surrounding the Rayleigh-step gas bear- 
ing by the high-pressure ambient, results in increased load capability. The 
sealing occurs across the dam which is a narrow annulus of lov clearance 
exposed to high pressure at its interior circumferential groove ind to lower 
pressure at its outboard end. The shaded regions on Figure 2-2 indicate 
depressions from grooves and Rayleigh-step pockets. Figure 2-3 is a photo- 
graph of a carbon ring seal set with embedded thermocouples for temperauure 
measurement . 

2.2 Test Results of the 50-mm Rayle igh-Step Helium Buffer Seal 

Four helium seal sets and three runners were tested during the coursr of the 
program. All were subjected to steady-state tests while two sets additionally 
underwent high acceleration rate t^sts. The tests resulted in 613 min of 
cumulative running time and 90 high acceleration rate starts. Table 2-1* 
provides a brief summary of the tests performed. 

The seals performed successfully over a broad range of conditions, including 
the full design pressure of 1379 kPa absolute (200 psia). The design speed of 
7330 rad/s (70,000 r/min) was not achieved. The first three seal tests were 
limited to about 5235 rad/s (50,000 r/min) due to a dynamics problem in tbe 
test rig. After modifications were made to correct the problem, a fourth seal 
test was run. 

Seal set No. 2 completed the test program to the maximum speed capability of 
the rig without damage, including 50 high acceleration start-ups. 



*Tables are presented consecutively, beginning on page 2-41, 



2-2 



The failure of seal f3t>t No. 1 waii dirtM-rly at; t ri butiablo r o lar^,e ampL i r.udf! 
motions of the seal runner. The Gstimated amplitude wa.s 0.038 to 0.(!'j8 mm 
(1.5 to 2 mils) peak to peak which was well beyond the acceptable ranp^e. Seal 
set No. 3 failed because of contamination in t'le helium supply system that 
infiltrated the ring clearance. 

Testing of seal set No. 4 was accomplished subsequent to modification made to 

the rig to achieve full speed. The changes clearly had a beneficial effect 

and allowed operation to proceed to 5968 rad/s (57,000 r/min) with mucn lower 

vibration levels. The modification, however, resulted in a higher temperature 

seal environment. Seal failure resulted by consumption of available clearance 

at a speed of 5968 rad/s (57,000 r/n.in). The rig did not show evidence of 

axcessive vibrations. The indications were that full speed operation of 

7330 rad/s (70,000 r/min) could be achieved, and seal failure would not ha./e * 

occu'-red if larger installed clearances were incorporated. 

Figure 2-4 indicates the steady-state data points for seal set No= 2, as a 
function of speed and helium supply pressure superimposed upon a theoL-etical 
operating range map. Analytically, there are two limitations on performance: 

1. Insufficient hydrodynamic forces to overcome friction forces - a 
low-speed, high-pressure constra..it identified by the high friction 
region on the figure 

2. Insufficient friction to counteract inertia forces - a high-speed, 
low-pressure constraint identified as the low friction region on the 
figure. 

Figure 2-4 shows an operating range map for the 50-mm seal that accounts for 
all constraints. If the pressure follows a speed squared relationship to a 
maximum of 1379 kPa (200 psia) at 7330 rad/s (70,000 r/min), it was expecte. ' 
that seal performance would be satisfactory. The majority of the data points 
fall into the high friction region where friction forces on the seal ring 
exceed hydrodynamic fluid-film forces. Most of the testing revealed that the 
seal ring remained stationary and shaft movement was contained inside the 



2-3 



clearance rof»ion. Any movement oi t h»> rin^» was . o anoilicr litatiunary po'jition 
to aceommodate variation in ahaft eccentricity at the ;jeal rin^ location. 

Figure 2-5 ahows flow data veraurj preogure drop tor the outboard neal ring on 
fjeal set No. 2 at various operating i^peeda. The solid linea reprofjent theore- 
tically predicted flows at various radial clearance;). Theoretical flow 1 1; 
independent of speed. Speed increapes cause clearance reduction because of 
centrifugal runner growth and thermal expanaionj-. Consequently, flow gener- 
ally decreased with speed. 

The moat accurate film thickness measurements were obtained from the outboard 
seal ring of seal set No. 4. Four probes were embedded in this ring and 
directly measured the film thickness between the seal ring and the shaft. 
Flow versus pressure drop data are shown on Figures 2-6 and 2-7. The resulta 
indicate that measured clearances are Lower than theoretical predictions for 
equal flows and pressures. However, there are indications of bypass flow 
between the seal rings and housing that could account for some of the differ- 
ences noted. 

2.3 C onclu sions and Re comm endat ion s for the bO-mm Rajrl^eig^r^te^ ^^^[ 

The general conclusions drawn from the 50-mm Rayleigh-step floating-ring 
helium buffer seal program are as follows: 

* The program results confirm Che capability of the seals to perform well in 
cryogenic turbomachines . Seal set Mo. 2 went through its complete test 
program unscathed and could be readi" • inserted into a machine and rerun. 
Eight steady-state runs were made resulting in a accumulative test time of 
261 min. Also, a total of 50 fast starts were completed with average accel- 
eration rates of over 152.4 m/s (500 ft/s ). Although full design inter- 
face speed of 183 m/s (600 ft/s) was not achieved, the seals performed very 
satisfactorily up to an interface speed of approximately 144 m/s (472 ft/s) 
which is 1.53 times faster than previously reported results for cryogenic 
applications [5]. In addition, full pressure drop of 1379 kPa absolute 
(200 psia) was accomplished up to a speed of 126 m/s (412 ft/s). There was 



2-4 



9 



fivery indication that a ^»r(jat.nr machinod cloaranco tor noiil 'id No. ^^ would 
have permitted fulL speed, tull ptannure operation. 

• Leakage flow ranges per seal ring averaged between 0.001 and 0.002 kg/s 
(13 to 25 sefm) and were not sij^ni f: ieantly affected by pro'jfjure drop and 
speed. The most sensitive parameter in clearance. Examination of Figure 
2-7 reveils that the measured flow varies approximately as the 2.5 power of 
clearance at a constant pressure drop. Figure 2-8 shows a maximum flow 
envelope for each seal set as a function of pressure drop. Each flow curve 
represent?, the higher of the two rings (inboard or outboard) flows at a 
given operating point. The highest flow recorded was 0.0026 kg/s (33 scfm). 
However, typical flows ranged between 0.001 and 0.002 kg/s (13 and 25 scfm). 
Figure 2-9 shows flow data for a single speed of 4712 rad/s (45,000 r/min). 
While it was not the highest speed tested, it is the highest speed at which a 
wide range of supply pressures were applied to onch seal set. Seal !iets No. 
2 and No. 3 had maximum pressL-re dropy of 1365 tcPa (198 psia) and 1250 kPa 
(181 psia). Maximum flow levels are 0.0019 kg/s (24 scfm). 

• Two of the three seals that failed by high-speed contact did so because of 
external factors unrelated to seal design and performance. C3ne failure 
occurred because of excessive shaft vibrations and another by a contaminated 
hel ium supply. 

• The seal design process requires integration with the machine environment. 
Thermal expansions and contractions and centrifugal growth of the runner are 
important considerations To correct test rig vibrations, changes were made 
that altered the seal environment. The program did not permit detail 
studies to be made to determine the effects of the environmental changi'i on 
seal set No. 4. As a result, the seals were insteil led with insufficient 
clearance. More liberal clearances would have permitted successful full 
speed operation. 

• The material combination (P-5N carbon graphite rings versus tungsten 
carbide coated runner) was not tolerant of high-speed rubs. The runner, in 
particular, performed badly. It retained carbon transfer, closing up seal 
clearances. Localized heating caused upset and flaking of the tungsten 



2-5 



carbide coating. Large nurtaco. ,irna nibfi alfio caufjcd "wipt'-out" ot the 
Rayleigh steps machined into Ene carbon rings. Investigation ot alternate 
material combinations was beyond the scope oi the program, but more suitable 
materials should be a future consideration. 

• The analytical tools used iii the design process were very ef-fi'ctive in 
predicting fluid-film performance and dynamic response. Initially, leakage 
was computed using viscous flow theory which proved to be difficult to 
correlate with experimental results. It was necessary to modify the flow 
theory to include entrance and film inertia losses (see Appendix A). This 
improvef'" correlation and interpretation, however, discrepancies stillexist . 
For Che same flow and pressure, theoretical clearances are larger than meas- 
ured by a factor of approximately 1.6. This is not a large factor consider- 
ing t-hp sensitivity uf flow with film ihickness. There was also evidence of 
bypass leakage along the end walls that was not accounted for by theory. 
However, further effort remains to explore in depth the reasons for the 
variations. 

• Dynamic response of the rings did not appear to be a serious problem. Anal- 
ysis indicated that the rings should have low mass to dynamically track 
runner axcursions. Consequently, they were made of carbon without composite 
metallic rings on the outer circumferences. In most instances, the rings 
were maintained in a static position. Test ri: speed limitations caused 
most test points to be in the high friction region of the operating range 
map (Figure 2-4), but there appeared to be sufficient hydrodynamic capabil- 
ity to move the rings when necessary. Dynamic response of the seal rings 
probably would have been more active at full design speed. 

"eal motions did arise on several occasions. The principal occurrence was 
with t.ie inboard ring of the seal set No. 3. The seal ring developed an 
in-phase, generally elliptical orbit. The orbit was observed at all three 
test speeds of 3665, 4188, and 4712 rad/s (35,000, 40,000, and 45,000 r/min) 
and became larger in amplitude as the supply pressure was increased. The 
maxiumm orbit diameter was approximately 0.025 to 0.030 mm (O.OOl to 
0.0012 in.). The runner orbit was approximately O.OIO to 0.013 mm 
(0.0004 to 0.0005 in.) so that some amplification was taking place. This 



2-6 



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rin>', did not tail nor make (unt.irt wi t fi t lit' runner and attt'sr t n t hi> i'x('p1 = 
lent dynamic qualit.ioi) ot Iho 'leal liyfitom. It, i ;] 'jpeculat.od that a fluid 
film existed between the rings and end wallo, minimizing', triccional re;,ist = 
ance of the ringrj to ahat t-induced motions. 

2.4 Results and Conclusiins tor the 20-mm Rayloigh-Step Helium Butler Seal 

For this seal, theoretical and desi^fjH investigations were made and the follow- 
ing is a summary of the significant results and conclusions: 

The maximum test rig speed is 10,472 rad/s (1(10,000 r/min). This linits the 
surface speed of a 20-mm diameter shaft to 104.5 m/s (343 ft/s). Thus, it 
wuld not be possible to test the ?0-mm tieai at the specified condition of 
lb3 m/'s (600 ft/s). Surface speed limitations also restrict fluid-film 
hydrodynamic force generation necessary to overcome seal ring friction. It 
was determined that the maximum buffer fluid pressure at 10,472 rad/s 
(IQQ^QQQ r/min) would be 689 kPa (100 psia). Figure 2-10 shows the operat- 
ing range map for the 20-mm seal, indicating the high and low friction 
regions that would cause difficulty. The operating range is narrower than 
for the 50-mm design. Also indicated on the curve is the speed squared 
relationship with pressure to be used in test. Regions of low friction are 
encountered as the shaft comes i'" to speed where excessive ring motions 
would occur if operation was sustained. However, if the seal system does 
not dwell at these conditions, it could pass through them without endanger- 
ing seal operation. 

• Centrifugal expansion for the 20-mm seal runner is negligible because the 
runner is integral with the small diameter shaft. Thermal contractions, 
however, require very close installation clearances of 0.0102 to 0.0229 mm 
(0.0004 to 0.0009 in.) diametral clearance. Miting the rings directly 
against the shaft is desirable since it removes the uncertainties of runner 
growth and distortions. Material combinations should be such that the shaft 
survive high-speed rubs. 

* As with the 50-mm seals, low mass and all carbon rings are required for 
dynamic tracking. 



2-7 



2.^ ^J^^J5£i£li£" and PrincipLo ot Operation ot the Spi raL-GroovG LOX SeaJ ;3 

The spiral-groove face aeal is a prime candidate for application to LOX turbo- 
pumpa. It is a fluid-film seal that can effectively inhibit laakage and avoid 
rubbing contact that could cause catastrophic explosion failure in a LOX envi- 
ronment. As described in Section 1.0, the function of the LOX seal is to aid 
in preventing leakage of LUX from the pump end of the machine. 

Originally, MTI examined a conventional type of spiral-groove seal that was 
labeled the straight-through design. The spiral grooves extended to the 
outside diameter and the fluid - 's pumped inward to a dam region at the inte- 
rior ID of the seal. Although excellent performance characteristics were 
prediccfid, the straight-through designs were ult'mately abandoned because of 
the probability of vaporization in the flow path. 

The pressure-balanced concept selected was conceived by NASA/LeRC p. d recom- 
mended for the LOX turbopump application because it obviated vaporization 
problems and reduced net axial loading. 

2.5.1 Special Considerations for LOX 

2.5.1.1 Flui d Vaporization . Because of the many restrictions the fluid pass- 
es through before it reaches the seal compartment, it enters relatively hot. 
However, since it is at high pressure, it is still a liquid. When passing 
through the seal, it is further heated by viscous shear and simultaneously the 
pressure drops to the seal exhaust ambient. Thus, upon discharging from the 
seal interface, it is in a mixed vapor state. At ambient pressure, the satu- 
ration temperatures of LOX is -183°C (-297. 4°F); the discharge temperature in 
the seal interface can be significantly higher and is generally around -115°C 
(-175 F). Therefore, vaporization is likely. Vaporization in the 
spiral-groove leakage path can block the through-t low which assists in carry- 
ing away the heat generated in the interface. This leads to further vaporiza- 
tion, reduction in fluid viscosity, and consequently load capacity. 
Ultimately, the seal fails due to overload and overheating. Therefore, an 
adequate seal design must handle vaporization without seriously jeopardizing 
performance. 



2-8 



2.5.1.2 Fluid T ur bulence. The criterion for turbulence i ;> that the ratio ot 
inertia to viscous fluid forces, which is the commonly applied Reynolds 
number, be greater than 1000 [6]. The Reynolds number is a dimensionless 
quantity defined as follows: 

Re = pUoh/M 



where: 



p = fluid density 

Uq = surface velocity 

h = film thickness 

U = absolute viscosity 

The characteristics of LOX are such that the mass density approaches that of 
water, while the viscosity is close to that of a gas. From the definition of 
the Reynolds number, this combination of fluid properties is conducive to the 
promotion of turbulence. The load on the seal would have to be great enough so 
that the operating filii thickness was only 6.1 x 10~ m (24 x' 10~ in.) for 
the flow to be' laminar. This film thickness is too small to be practical and 
would produce heat generation that would vaporize the fluid and cause failure 
by contact and overheating. The final seal design operated with film thick- 
nesses near 2. 54 x 10 m (0.001 in.) with resulting Reynolds numbers in the 
vicinity of 42,500. The effects of turbulence upon fluid-film performance is 
the same as if the seal were operating in the laminar regime with a greatly 
increased viscosity. It produces a greater load carrying capability and high- 
er viscous power losses than a laminar operating seal. With respect to the 
spiral-groove geometry, turbulence results in much deeper and wider grooves 
than a laminar counterpart, so that sufficient fluid of apparent high viscosi- 
ty could be pumped through the large clearance. 

2.5.1.3 Fluid Inertia . Because of the large operating film thickness and 
high-pressure gradients, the flow restriction through the seal is not entirely 
viscous. Inertia forces are another significant consideration. Inertia 
produces steep pressure drops at flow restrictions, or sudden contractions in 
the path of the flow. At the sudden conti. act ions , pressure head is converted 



2-9 



to veLocity head. Such contractions occur at the spiral -groova dam upon tluid 
entrance in>.o the seal land, and also upon fluid entrance into the 
spiral-groove seal from the high-pressure fluid at the oucer periphery. 
Another inertia influence is centrifugal forces acting on the fluid. These 
were not considered because of the complexity of the analyses; their general 
effect is to retard flow. It was estimated thev would not have more than a 10% 
influence on the predicted rates. 

2.5.1.4 Two- Phase Flow. Once the fluid vaporizes, it enters the mixed flow 
regime and a considerable amount o£ energy is required to boil the liquid. 
When it is in the mixed regime, the performance of the seal is not precisely 
known although it was speculated that leakage would be close to that of a pure 
liquid. Because of this, most of the seal interface is in the spiral-groove 
region where the fluid is at high pressure and in a purely liquid state. It is 
only at the latter end of the seal dam where the pressure is low that vaporiza- 
tion generates. Two-phase flow may be a significant phenomenon for future LOX 
seals and is an area where further effort should be devoted. 

2.5.2 Pressure-Balanc ed S piral-Groove Seal 

The pressure-balanced concept is schematically pictured on Figure 2-11. The 
spiral-groove region is fed by an interior groove that communicates through 
passages (drilled holes) with the pressurized fluid to be sealed. In this 
instance, the term pressure-balance implies that the high-pressure fluid to be 
sealed resides at both the inside and outside perimeters of the spiral-groove 
region. The grooves pump outward to a dam region on the outer periphery of the 
seal. Leakage flows from the interior groove inward through a sealing land to 
the ambient low-pressure region. 

A principal advantage of this configuration over the straight-inflow design is 
that the spiral-groove pumping circuit operates independently of the leakage 
circuit. Thus, if the fluid flashes or vaporizes in the leakage dam inter- 
face, it will not affect the pumping action of the spiral-groove and the abil- 
ity of the faces to maintain separation. From a leakage point of view, 
flashing is beneficial since it significantly reduces mass flow. The one 



2-10 



major problem with Che preasure-balancec' concept is Che posgibilicy of over- 
heacing. 

Consider a concrol volume surrounding Che seal, as represented by the dashed 
lines on Figure 2-11. What leaves and enters the control volume is leakage 
flow. However, internal to the control volume, there is a significant amounC 
of viscous shear. If we presume that all the heat generated by viscous shear 
is carried away by Che leakage flow, Chen the temperature rise inside the 
control volume can become very high. However, if external cooling flow is 
circulated through the control volume, it can carry away the heat generated by 
viscous friction and prevent excessive temperature rises. Fortunately, 
excess flow is forced into the seal cavity of many LOX turbopumps and recircu- 
lated into the impeller inlet (refer to Figure 1-1). This circulating flow .^ 
can prevent overheating. 1 

A typical pressure distribution across the face of the pressure-balanced 
concept is shown on Figure 2-12. The pressure is normalized with respect to 
the high pressure being sealed on the outer periphery. The radial distance or 

span is indicated nondimensional ly along the abscissa. The origin in Che ^ 

I 

radial direccion is at the inleC of Che spiral-groove region. The pressure -i 

1 

increases from Che inCerior of Che grooved region ouCboard uncil Che 

spiral-groove dam Is reached. Then, Chere is a sudden inertia pressure drop 1 

at the entrance to the dam, followed by a linear pressure drop through Che dam ;J 

Co Che ouCer periphery of Che seal. From Che origin ac Che incerior of Che 
spiral-groove region, proceeding inward, chere is a corisCanC pressure in Che 
grooved annulus followed by a significani; pressure drop aC Che enCrance Co Che 
sealing dam and Chen a linear drop across Che sealing dam Co ambienC pressure. 
The inerCia pressure drop is higlT across Che sealing dam because of Che high- 
pressure gradienc across ic. 

2.6 50-mm Spiral-Groove Seal 

2.6.1 Groove GeomeCry 

The general groove geomeCry is shown on Figure 2-13. Note ChaC deep and wide 
grooves are necessary Co pump the highly turbulent fluid because of its high 



2-11 



effective viscosity. Table 2-2 defines the principal nominal dimensions of 
the pressure-balanced design. 

The original MTI design layout of the 50-mm pressure-balanced seal is shown on 
Figure 2-14. In the test rig, the seal will be installed in a back-to-back 
configuration as r^hown on the bottom left corner of the drawing. This was 
done to eliminate excessive thrust loading on the test rig thrust bearing. 
The outboard seal is the test seal, while the inboard seal is the thrust 
balancing seal. The nonrotating member of the spiral-groove seal is made from 
carbon graphite (P-5N) and contains the interior high-pressure inlet annulus 
and feed holes. 

The seal rings do not have metal shrouds or interface pieces. The intent is to 
maintain the mass of the seal ring members as low as possible for improved 
dynamic response. Figure 2-15 is a photograph of one face seal and a mating 
ring. Because of the back-to-back installation, the mating ring incorporates 
grooving on both sides, but of opposite hand. 

2.6. 2 Summary of Pe rfor mance 

The principal advantage of the pressure-balanced design is that it separates 
the spiral-groove and leakage circuits, and reduces the consequences of fluid 
vaporization in the seal region interface. 

Disadvantages of the pressure-balanced design are high power loss including 
viscous friction and windage of the rotating matini; ring, relatively large 
collar diameter, poor low-speed performance, and high lift-off speed. Also, 
without external cooling or recirculation of LOX inlet yupply, there is a good 
possibility the seal will overheat. A summary of calculated performance 
results at design operating conditions is luviicated on Table 2-3. The table 
provides information for both the cooled and unccoled inlet conditions. 

The factors that influence safe operation of pressure-balanced seals are 
film-thickness, fluid temperature rise, and dynamic response or ability of the 
seal ring to follow excursions of the rotating collar. Based on the analysis, 
an acceptable operating range was established. This is shown in Figure 2-16. 



2-12 



At low rotational speeds, low film thickness and excessive temperatures are 
limiting factors. At lower pressures and high speeds, the limiting factor is 
dynamic response. The safe operating range lies within these boundaries. 
Note that a pressure versus speed squared relationship will produce safe 
start-up operation. The relatively large operating clearance provides a 
greater tolerance to distortions and dynamic response than laminar seals that 
typically operate in the film thickness range of 0.0025 to 0.0051 mm (0.0001 
to 0.0002 in.) 

The effects of turbulence are increased film thickness and power consumption. 
Geometrically, th? seal requires deep and wide pumping grooves. The magnitude 
of the turbulence is large, producing Reynolds numbers in the 40,000 to 50,000 
range. 

2.7 20-mm Spiral-Groove Seal 

2.7.1 Geometry 

Figure 2-17 is a layout drawing of the 20-mm pressure-balanced design. The 
concept is very similar to the 50-mm design except it is a scaled down version 
consistent with the size reduction, but optimized to operate at similar 
surface speeds of 183 m/s (600 ft/s). Table 2-4 indicates the pertinent nomi- 
nal dimensions of the 20-mm design; the spiral-groove geometry is also indi- 
cated on Figure 2-17. 

2.7.2 Summary of Performance 

Performance of the 20-mm seal is summarized on Table 2-5 and Figure 2-18. The 
table presents performance at the design point condition, while the figure is 
an operating range map. The map indicates low-speed, high-pressure operation 
is to be avoided because of marginal film thickness and high temperature rise 
of the fluid. At high-speed, low-pressure operation, dynamic response prob- 
lems are encountered. A pressure increase proportional to the squa'-e of the 
operating spned produces safe operation. 



2-13 



2.8 Test Rig 

2.8.1 Design Phil osophy 

The general design of the test rig was guided by several fundamental princi- 
pals generated at the outset of the program. 

1. The rig was to have the capability to test both 50- and 20-mm seals. The 
50-mm seals were to be tested at a maximum speed of 7,330 rad/s 
(70,000 r/min); the 20-mm iieals were to be tested at a maximum speed of 
10,472 rad/s (100,000 r/min). 

2. Different shafts could be installed for the 50- and 20-mm seal tests, but 
the same set of bearings were to be employed. A journal size of 30 mm was 
selected as a compromise for testing both size seals and for providing 
acceptable rotordynamic response and bearing power losses. 

3. Hydrostatic fluid-film bearings were to be employed to enable friction- 
free* start/stops, whirl-free operation, good damping qualities, and free- 
dom of adjustment through flow and restrictor element alterations. 

2.8.2 General Configuration 

A cross section of Che test rig is shown on Figure 2-19. The right-hand 
portion of the rig is the drive where the nitrogen turbine is located. The 
central portion is the bearing region where the journal and thrust bearings 
are located. The left end of the rig is the test seal section; the SO-mm heli- 
um buffer seal is shown installed. 

Locating the thrust bearing in the center of the rotor avoids excessive over- 
hang at either end and provides for a more uniform distribution of mass along 
the rot'r. This arrangement alleviates rotordynamic difficulties due to large 
overhung masses which would oc'jur if the test seals and thrust bearing were 
mounted in tandem at one end of the rotor. The helium buffers are installed in 
a back-to-back configuration, and mate against a common runner. 



2-14 



1 
I 
J 
I 
I 
I 

I 

I 
\ 
i 
I 
I 
[ 
I 
I 
1 
i 



The 30-mm shaft journal diauieter providoa rjuffJcient stiftneos to be below the 
bending critical speed, and prevents excessive bearing and windage power loss- 
es ^or operation at 7330 rad/s (70,000 r/min). At the turbine end of the 
shaft, a heat dam is located between the turbine wheel and shaft. This dam 
prevents high temperature at the turbine wheel from conducting heat into the 
cold shaft regions. The outside periphery of the heat dam is machined with a 
labyrinth that provides one half of a buffer seal that prevents turbine gas 
from entering the bearing region. At the seal end of the shaft, the helium 
seal runner is secured to the shaft by a spring sleeve that is pressed onto the 
shaft. This compensates for bore growth of the runner. Figure 2-20 shows 
disassembled components of the test rig and Figure 2-21 is a photograph of an 
assembled rotor. Details concerning the rig design are presented in 
Section 6.0. 

Because of problems uncovered during the first three tests, three modifica- 
tions were made prior to the fourth seal test. Perhaps the mcst significant 
alteration was the addition of a small labyrinth seal between the inboard 
helium seal ring and the nearby journal bearing as shown in Figure 2-22. 
Helium leaking from the seal and an additional flow of helium entering through 
a new external port, were maintained at a slightly higher than bearing drain 
pressure, preventing LN2 from flowing through the new labyrinth seal and bath- 
ing the inboard end of the runner. This was done to reduce the heat generation 
from windage losses that was causing vaporization in the adjacent bearing 
film, reducing bearing effective stiffness and damping and resulting in unac- 
ceptable rotor response at speeds below design speed. The insertion of the 
labyrinth seal insured that only gaseous helium wf.uld run against the seal 
runner with significantly lower windage losses than Chat produced by LN2 . 

2.8.3 Summary of Test Rig Performance 

In general, the test rig performe-. well with the principal exception of not 
achieving full speed. As mentioned above, this problem was traced to vapori- 
zation in the seal end journal bearing reducing ity stiffness and damping 
characteristics which resulted in high vibration levels at approximately 
5236 rad/s (50,000 r/min). Corrections were made to the rig to eliminate this 



2-15 



problem, but a seal tailure at 5969 rad/a (57,000 r/min) prevented tvirther 
speed increases. 

The testing program surfaced several other characteristics of the test rig 
that are significant and peculiar to cryogenic testing. These are mentionable 
because they have general applicability to cryogenic test rigs and are impor- 
tant CO understand for future applications of the rig. 

1. Initial Cooldown. The hydrostatic bearings are energized early in the 
start-up process in order to float the shaft prior to rotation. Initial- 
ly, the rig is at ambient temperature and the cryogenic fluid supply to 
the bearings (LN2 or LOX) is in a gaseous state. Since there are deep 
recesses in the bearings, the bearing rotor system is prone to pneumatic 
hammer until the cryogenic fluid is liquefied. It was necessary to manu- 
ally hold the shaft from vibrating by forcing it axially against the 
thrust bearing surface during the chill-down period ("^ZO-min time span). 
For LN2 testing, this requirement presented an inconvenience but no seri- 
ous problems. For LOX testing, however, personnel are not permitted in 
the vicinity of the test rig, and a remote holding device is required. 
For future hydrostatic bearing designs, the recess volume should be as 
small as practicable to avoid pneumatic hammer at startup. 

2. Recess Pressure M easurement. Measurement of recess pressures provides 
information on the state of health of the rotor-bearing system while oper- 
ating. Unfortunately, the measurement system introduced problems and 
ultimately had to be abandoned. What occurred was vaporization of the 
cryogen in the pressure sensor lines emanating from the rig to the pres- 
sure transducers. This deteriorated bearing stiffness because the fluid 
column stiffness in the instrument line is significantly less when vapor 
is present. Also, heat is transferred from the line fluid into the recess 
promoting bubbling. It was necessary to plug ■' . pressure tap lines at 
the tester and operate witnout recess pressure information. 

3. Isolation of the seal test area from the seal end journal bearing was 
required to prevent heat transfer into the bearing and causing vaporiza- 
tion and bubbling of the cryogen in the bearing. The installation of a 



2-16 



labyrinth seal between the bearing and seal eomparrment and an additional 
source of helium buffer fluid prevented ingress of LOX to the bearing side 
surface of the seal runner. A substant i.:*! reduction in windage heat 
generation was accomplished, which effectively eliminated excessive heat 
transfer into the bearing from the seal compartment. 

4. Nitrogen was not an acceptable buffer flui d at the turbine end buffered 
labyrinth seal because it liquefied rapidly. It was necessary to use 
helium as the buffer fluid. 

2.9 j.n3trumentat ion 

There is one area regarding instrumentation that is worthy of mentioning in 
this summary of significant results. The original method of measuring seal 
film thickness was to trace the movement of the seal rings and runner sepa- 
rately and obtain clearance by electronically subtracting signals. It was 
originally determined that embedding probes directly into the seal ring would 
impose restraint by cables that would prevent dynamic tracking. It was subse- 
quently discovered that a very thin coaxial cable (0.76 mm in diameter) 
existed, that would not impose excessive restraint on the seal rings. It was 
decided to make probes and embed them directly into one of the seal rings for 
test No. 4. Direct measurement of film thickness rather than the differential 
approach would produce more accurate measurement and errors due to large 
temperature gradients in the housing could be avoided. The probes performed 
very well . 

Figure 2-23 shows details of the probe installation and Figure 2-24 is a 
photograph of the instrumented ring. The two original probes observing the 
seal runner were kept and used to monitor runner motion. The embedded probes 
were not subject to error due to thermal distortions of the seal housings. 



2-17 



Helium Buffer Fluid 




i 



OJ 



Ff = 
Fr = 



Pressure Balance Force 
Normal Contact Force 
Friction Force 
Hydrodynamic Force 
Hydraulic Pressure Closing Force 



N 




N 



Fig. 2-1 Floating-Ring Seal Schematic 



■3022 



r' 



nr HZ 'HI nz cz 



C2 cz nz nr CD 



I 
I 



Circumferential 
Groove 



Pa 



Seal Dam 




Groove Depth = 0.229 mm 



Axial 
Groove 




Po 



} 



■Pocket 
(Depth 
0.0254 mm) 



Section A-A 



832437 



Fig. 2-2 Developed View of 50-mm Rayleigh-Step Pad 



2-19 



Xinvnb MO'-.! to 

81 aOVd TVNIDIMO 




Ll 



u 



l\ 



Pig. 2-3 50-mm Rayleigh-Step Floating Rin" , 



■M 



r 1 



. 1 



OPi" 



ur^M- P^^^^ 'v 



OK '' 



\;..V- ' 



2-20 







1500 






1400 






1300 






1200 




« 






D 

■m 


1100 




• 
a. 


1000 




^ 






Ui 

cr 


900 
800 




UJ 

Of 

a. 


700 


1 


>- 






o. 
a 


600 
500 



4C0 
300 
200 
100 





1000 



2000 



3000 4000 5000 

SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Fig. 2-4 Operating Map for SO-na Helium Buffer Seal - Seal Set No. 2 



►r 



^^1 



.ao7r 



.006 - 



.005 



u .004- 

a 

3i 



I 



Ezzizni: 



1*3.4 

s/.t 

l.ISi 

••• 

l.« 

l.ii 
4.0t3 



OOWNSfREAN PftCSStAC (VRs) 
UPS1RCAM HMPCRAIUM <(t«g C) 
SCAL LCNG1H (M) 
SCAL DIAN£ICR (m) 
ECaNIRICIlY RAIIO 
DISCHARGE COCFFICKNT 
VISCOSITY (Pa-s) 
AOIAIATIC COEFFICIENI ( 
MOLECULAR WEIGN1 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 2 
OUTIOARO SEAL 



> 




400 500 600 700 BOO 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 2-5 Outboard Seal Flow, 103 kPa Drain Pressure - Seal Set 



No. 



nz cz 



MO 



, .j: cr: rr [~ K C- 



I 
I 
i 
I 

L 
C 
L 
C 
I 
I 

i: 
i: 

t 

c 




100 



200 300 400 

PRESSURE DROP ACROSS SEAL (KPt) 



too 



Fig. 2-6 Zero Speed versus Pressure Drop - Seal Set No. 4 




100 



200 300 400 

PRESSURE DROP ACROSS SEAL (KPa) 



500 



«0« 



Fig. 2-7 Flow versus Pressure Drop - Seal Set No. 4 

2-23 



0.003 



I 



0.002 



(0 

at 
o 



0.001 




o 


Seal Set No. 1 


A 


Seal Set No. 2 





Seal Set No. 3 


D 


Seal Set No. 4 




Speeds: 3141 rad/s to 
5968 rad/s 



- Design 
Pressure 



J I I i i_ 



200 400 600 800 1000 1200 1400 



Pressure Drop Across Seal (kPa) 



Kig. J-d St. 1 1 l.cak.i^L- Kiwi: lope Utila al Varying; Spci-ds 



851 1S5 



□ [n: d c: nil nz Cj 



r-^c:nia:nzc::'::2czczc:^c2 



1 W 9 • 1 



0.003^ 



I 



0.002 



s 
I 



0.001 




A. 



J I L 



J I I I I I L 



o 


Seal Set No. 1 


A 


Seal Set No. 2 


C 


Seal Set No. 3 


D 


Seal Set No. 4 




Speed: 4712 rad/s 



Design 
Pressure 



200 400 600 800 1000 1200 1400 

Pressure Drop Across Seal (kPa) 

Ki^. J-4 SimI l.i-.ika^f Kiivil opf D.ii.i .il Cuii.si .ml Spttd 



851 156 



G 



700 r 



600- 



£ 500 

« 

a 

Q. 

« 400 

3 



i 300 



"5 



5 200 

3 
CD 



100 



• 




100 


- 


. ^ 


V 








^^ 




- 




80 


>// 


D 


K 






y/^<^ 


V' 




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a 


60 


High Friction y/^ Ly<^ 


— " 1 




3 

1 

Q. 






r^/ 


U 




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u. 


40 


sJ^ 


/*^^ P a N^ 


M 


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^ 




y / y 


/ 




- 


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OQ 


20 


_ Low Friction 


1 

1 








Q 




1 1 1 1 1 







20 40 60 80 100 






Shaft Speed (ipm x 10') 


n 








1 1 1 I ■ i ■ 1 1 1 1 i 


a 




1000 3000 5000 7000 9000 11000 






Shaft Speed (rad/s) 








—J 


Fig 


2-10 Operating Range Map for 20-inm Helium Buffer Seal 


D 








fi 








^^^ 






83107-1 










2-26 




D 



Housing 



\ 



Spiral Grooves 

Seal Ring 
Secondary Seal 



Load 
Spring 



Leakage 




Spiral-Groove 
Mating Ring H 



~1 Cooling 




Circulating 



Carrier 



Leakage 



83012-1 



Fig. 2-11 Pressure-Balanced, Outward Pumping, Spiral-Groove Concept 



2-27 



Iv> 



00 



I 

o 

Q. 



1.4,- 



1.2 



1.0 



0.8 



Q. 

, 0.6 



0.4 



0.2 




Feed Groove 




± 



± 



± 



1 



-0.8 -0.6 -0.4 



-0.2 0.2 0.4 

(R -Rl) / (RO - Rl) 

Fi^. 2-12 Seal Face Pressure Profile 



6 0.8 1.0 



83109 



[r:Lri^Lr:iir[r:cjCrfZ!C^r:-iL^cni 



2.54-mm DIa. — 16 Holes 

Equally Spaced on 

33.78 Radius 



h» 







30.48-mm I. Rad. 
Seal Land 

32.51 -mm I. Rad 
Seal Groove 

35.05-mm O. Rad 
Seal Groove 



35.05-mm I. Rad. 
Rotor Spiral Grooves 

38.86-mm O. Rad. 
Rotor Spiral Grooves 

40.1 3-mm O. Rad. 
Seal 



No. Oi Grooves = 15 
Groove Depth = 0.254 mm 
Land Width 



Groove Width 



= 0.45 



Fig. 2-13 Groove Geometry 



83114 



i. 



1 



— - \if \/t0o «is r 








Mtai ^rmu jrmmt wtmKt^ /kM4» .«& 

tUtrtSatattMAH^t »».■ »■<■ <«M • <«*<j 



fttt tft ■ . ««« 
Ml* \M -t-M* ^•••'r try^lt ^-^ ^r*mm 

. M. «4 Jr. /« « <» 













^1« f fc»i»« JS»r«« J«<«« ««aM«« 












fif 






T 



T 




j^J^ 






Mtwt Ifl 




Pig. 2-14 50-iin Pressure-Balanced LOX Seal 



•siaao 



rr:- Lij 



c- CZ dCZ. c. 



cz 



i:4 j-'/w uuA*.iiV. 




f i | ' l'i'| ' i'l'i'| i |'|i|i|'|i|iifii|i;im:- r i| 



lUOf .'•WiUI 



;j 



Pig. 2 -IS 50-am SpiraL-Croove Seal, Hating Ring, and Face Seal 



2-31 



rad/s X 103 



n: 
I 



M 
Q. 

W 
0) 
O) 

M 

o 
o 

< 

Q. 



Max. AP 



Safe Operating 
Range y 

/ 




/ 



/i 

I 
I 

I 

Operating | 
Speed ^ 

I 



Max. Rig 
Speed ^ 



Poor 

Dynamic 

Response 




30 40 50 

Rotational Speed, rpm x i03 



50 



4 



30 



20 



10 



(0 
CL 



80 



Fig. 2-16 Operating Hange Map for SO-min Spiral-€roove Seal 



sni9 



a rzi 



[^ 



• a rr r? c [i: c: Lz: [_' 3: cr- cz l^: cz 



wn n 



1 



i. 



1 




m» /( Str*^ C»m-»^ As»€m S-^ 0m. S9. 0'^m* 






— 1lf|/<IWXI5 I* 



t**M9 B*—^^^€ Mm i i ^^ > tf. 






















#ftf t ta»\»» Mtf»* i»**m ««•«•«• 




I ttmrmit (*•*'/■>« C|«i n -i< tS*i 



/■M !•«.«» #'lril . 

/T<l«« Iff _ 

. _ f IV Till Ait 1 11 
1 lit *»•• J_ 

MI 4 - "^ ^ «v4-*>MMMBr 



i^'jUi^iwi- 



EEiI 



Fig. 2-17 20 Pin Pressure-Balanced LOX Seal 



•sisao 



Max AP 



I 

5^ 



Q. 



« 
CO 

CO 
CO 

O 

u 

a 

Q. 
<1 




Maximum 
Speed 



20,000 40,000 60,000 

Rotational Speed (rpm) 



80.000 100.000 



J L 



2.000 4,000 6,000 8,000 

Rotational Speed (rad/s) 



10,000 



83020-1 



Pig. 2-18 Operating Range Nap for 20-ian Spiral-Groove Seal 



(n m m 



rz □ 



rri cr: cz r^ c: cl" 



_ I. 



m cr d! 



m n n n 



Capacitance 
Probes 



I 



Seal 
Housing 



Outboard 

Oram 
Cavity 



LN2 Journal 
Bearing 




O O 



Kig. _»-14 Test Rig Cross Soctiun 



■6n4S-1 



LaOyrinth 
Hot >ing 




Spacer 



LOX Fi 



Haiium Sail 
Runner 



Fig. 2-20 Test Rig Components 



ORIGINAL PAGE IS 
DE POOR QUALITY 



2-36 







ThrutI Coll«r 



Helium Sail nunner 



Fig. 2-21 Helium Seal Rotor 



2- 7 



G 



'S»< 



Low-Pressure 

Helium Supply 

Port 




Labyrinth 
Seal 



Fig. 2-2 hodified Test Rig Configuration - Seal Set No. 4 



851590 



2-38 



'J 

u 
u 

[J 

a 

a 
ti 
n 



ij 
ij 

!.l 

I] 

D 
D 



i -. *- 



I t I i 



0090 
0080 



(He«) 



Capacitance 

Probe 

(4 Places) 



I 



12 in ±Vi6 in 
(Typ) 




UT-13 Cable 
— (0 023 m O D ) 
(Maximum Flexibility) 




\ 
0004 
0002 



42% iron -Nickel Alloy 

(Matches Expansion Rate of 

Carbongraphite Seal Rir>g) 

Assemble As Shown Then Gnnd 
Probe Flush with Surtace 



Detail A 

(Scale: 10X) 
(Typ. 4 Places) 



Fig. 1-T\ Details ot Eiri>edded Prt>be Insial latiuii 



8S237 



■BClNAr; PAGE IS 
■ BOOR Oi'MJTY 



Embedded Capacitance Probes 




•Embedded Capacitance Probes 



Fig. 2-24 Instrumented Seal Ring with 
Embedded Capacitance Probes 



2-40 



TABLE 2-1 



SUMMARY OF TEST RESULTS 



Hardware 
Tested 


MaziDum Supply 
Pressure, kPa 
Absolute (psia) 


Haxinmm 

Speed 

rad/s (r/min) 


Cumulative No. o£ 
Run Time Fast 
(min) Starts 


Hardware Conditions 


SEAL SET NO. 1 
aunner No. 1 


1,179 (171) 


5,895 (56,300) 


96 





Outboard seal rubbed 
due to tester 
vibration. 

Outboard seal ring 
worn. Runner has 
surface cracks. 


SEAL SET NO. 2 
Runner No. 2 


1,482 (215) 


5,099 (48,700) 


261 


50 


No damage. 


SEAL SET NO. 3 
Runner No*. 2 


1,482 (215) 


4,963 (47,400) 


236 


40 


Outboard seal rubbRci 
due to contamination. 

Outboard seal ring 
worn. Runner has 
surface cracks. 


SEAL SET NO. 4 
Runner No. 3 


827 (120) 


5,968 (57,000) 


20 





Inboard seal rubbed 
due to loss of 
clearance. 

Inboard seal ring 
worn. Runner surface 
galled and worn. 



2-41 



TABL E 2-2 

DIMENSIONS OF PRESSURE-BALANCED, 50-MM , 
OUTWARD PUMPING, SPIRAL-GROOVE SEAL" 



Outside Groove Dam Radius: 40.13 mm (1.58 in.) 

Inside Groove Dam Radius: 38.86 mm (1.S3 in.) 

Inside Groove Radius: 35.05 mm (1.38 in.) 

Outside Seal Dam Radius: 32.51 mm (1.28 in.) 

Inside Seal Dam Radius: 30.48 mm (1.20 in.) 

Secondary Seal Radius: 30.99 mm (1.22 in.) 

Inside Ring Radius: 25.4 mm (1.00 in.) 

Seal Length: 20.3? mm (0.8 in.) 

Secondary Seal Position 

(Distance from Face): 15.75 mm (0.62 in.) 

Groove Depth: 0.254 mm (0.010 in.) 

Groove Angle: 11° 

Land Width/Groove Width: 0.45 



2-42 



I. 






TABLE 2-3 



STEADY-STATE PERFORMANCE OF 
50-MM SPIRAL-GROOVE SEAL 



Film Thickness, mm 
(in.) 

Axial Stiffness, N/mm x 10"^ 
(lb/in. X 10 ■^) 

Leakage, m /a x 10 
(in.Vs) 

Power Loss, kW (hp) 

At Grooves, °C (°F) 
AT Seal, "c (°f) 
Load, N (lb) 
N, rad/s (r/min) 



Pressure 


Pressure 


Balanced 


Balanced 


-Cooled- 


Uncooled 


0.0236 


0.0236 


(0.0009) 


(0.0009) 


81.1 


81.1 


(463) 


(463) 


29.8 


29.8 


(18.18) 


(18.18) 


9.47 (12.69) 


9.47 (12.69) 


-118 I 180.4) 


-104 (-155.2) 


21.44 (38.6) 


35.56 (64) 


2.83 (5.09) 


17.22 (31) 


10,608 (2,385) 


10,608 (2,385) 


5,864 (56,000) 


5,864 (56,000) 



P = 5.17 MPa (750 psig) 



2-43 



TABLE 1-j^ 

DIMENSIONS OF PRESSURE-BALANCED, 20-MM^ 
OUTWARD PUMPING, SPIRAL-GROOVE "seal 



Outside Groove Dam Radius: 25.91 mm (1.02 in.) 

Inside Groove Dam Radius: 24. 6A mm (0.97 in.) 

Inside Groove Radius: 20.8.3 mm (0.82 in.) 

Outside Seal Dam Radius: 18,29 mm (0.72 in.) 

Inside Seal Dam Radius: 16.26 mm (0.64 in.) 

Secondary Seal Radius: 17.53 mm (0.69 in.) 

Inside Ring Radius: 11.18 mm (0.44 in.) 

Seal Length: 20.32 mm (0.80 in.) 

Secondary Seal Position 

(Distance from Face): 15.75 mm (0.62 in.) 

Groove Depth: 0.203 mm (0.008 in.) 

Groove Angle: 11° 

Land Width/Groove Width: 0.45 



2-4' 



TABLK 2-5 

STEADY-STATE PERFORMANCE OF 20-MM 
SPIRAL-GROOVE sIaL 





Pressure 

Balanced 

Cooled 


Film Thickness, mm (mils) 


25.9 (1.02) 


AxiaL Stiffness, N/m x 10"^ 
(lb/in. X 10~5) 


'•.35 
(2.77) 


Leakage, m? / 3 x 10^ (in.^/s) 


1.9 (11.6) 


Power Loss, kW (hp) 


7.1 (9.5) 


At Grooves, "C (°F) 


25 (A5) 


At Seal, "C (°F) 


2.::2 (4) 


Load, N (lb) 


6383 (1435) 



P = 5.17 MPa (750 psi) 
N = 10,472 rad/s (100,000 r/min) 



2-45 



3 . TEST R ESULTS 

]L1 Introdiuetion 

Four helium seal sets and three runnera were tested durini?, the course ot the 
program. All were subjected to steady-state tests while two sets additionally 
underwent high acceleration rate teats. The tests resulted in 618 min ot 
cumulative running time and 90 high acceleration rate starts. (Refer to Table 
2-1 for a brief summary of the tests performed.) 

The following sections provide a detailed report of the results of the tour 
seal testa and their correlation with theoretical predictions. 

The test rig configuration for the first three seal tests followed the 
original design (see Figure 2-19). A 50-mm seal runner is secured to the left 
side of the rotor. The two seal rings fit back-to-back over the runner and are 
held apart by a spring. They are restrained on the right side by the seal 
housing and on the left side by a ring bolted to the housing. The seal rings 
are made of carbon graphite while the seal runner consists of Inconel 718 with 
a tungsten carbide coating on the outside diameter. 

Helium gas is supplied through two radial ports to the annular space between 
the seal rings. Part of the flow passes between the runner and left seal ring 
(outboard) exiting into the large drain cavity on the left side of the tester 
and from there into a separate drain line. The remainder of the helium flow 
passes between the runner and the right-hand seal ring (inboard) into a cham- 
ber between the seal housing and adjacent bearing housing. From the nearby 
journal bearing, LN2 also flows into the chamber and mixes with the helium. 
The mixture then flows through a passage under the journal bearing and out 
through one of several radial ports. The LN2 has a substantial cooling effect 
on the end of the runner and acts to increase the operating clearance of the 
inboard seal. 

Six capacitance probes were installed to measure the seal film thickness. 
Three of the probes are oriented horizontally while the other three are 
directed vertically. In the vertical group, two probes observe the back of 
the two seal rings. (Refer to Figure 2-19.) A third probe^ not shown, 



3-1 



observes the surface of the runner between the rings. The vertical film 
thickness is derived by electronically subtracting the output of the respec- 
tive seal probe from the commen runner probe. The horizontal components are 
similarly measured. 

Because of problems uncovered during the first three tests, three modifica- 
tions were made prior to the fourth leal test. Perhaps the most significant 
was the addition of a small labyrinth seal between the inboard helium seal 
ring and the nearby journal bearing as was shown in Figure 2-22. Helium leak- 
ing from the seal and ,;" additional flow of helium entering through a new 
external port were maintained at a slightly higher pressure than the bearing 
drain pressure. Thij prevented LN2 from flowing through the new labyrinth 
seal and bathing the inboard end of the runner. It also was done to reduce the 
heat generation from windage losses that was causing vaporization in the adja- 
cent bearing film, reducing bearing effective stiffness and damping and 
resulting in unacceptable rotor response. The insertion of the labyrinth seal 
insured that only gaseous helium would contact Che seal runner with signif- 
icantly lower windage losses than that produced by LN2. 

In place of the six-probe differential array, an outboard seal ring containing 
four embedded probes was installed permitting a direct measurement of film 
thickness during the fourth test. Figure 2-23 showed the instrumented ring. 
The two original probes observing the seal runner were kept and used to moni- 
tor runner motion. The embedded probes were not subject to error due to ther- 
mal distortions of the seal housings. 

A final change consisted of a new runner with electrolized surface to 
replace the tungsten carbide coating used in the original design. The rubs 
which occurred during the first and third seal tests resulted in partial 
delamination of the coating which contributed to tha failures. It was hoped 
tne electrolized runner would provide a better surface in the event of a rub. 

The pretest seal clearances are given in Table 3-1*. The overall accuracy of 
the measurements is approximately ±0.005 mm (±0.0002 in.). For the most part, 



*T8ble3 are presented consecutivply, beginning on page S—'+A. 



3-2 



Che pretest cLearancts of the tirgt three seal aeta conform to the defiign 
requirementa. The maximum discrepancy, found on the No. 1 outboard Qeal , waa 
0.002 mm (0.00008 in.), certainly within the range of meaaurement accuracy. 
The pretest clearances for the fourth seal aet are slightly larger Chan ch'^ 
original design requiremenCs. Thia waa conaidered desirable because of Che 
higher runner temperatures that were expected due to the installation of the 
new labyrinth seal. 

3 . j^ S t eadjT-S t at e J'e a^ Je s uU^ 
3 . 2 .J Seal Set No. 1 

Principal data was taken during three runs conducted over an 82-min period. 
Figures 3-1* and 3-2 provide time history plots of the test showing tester 
speed and helium supply pressure. Figure 3-3 gives an operating map with the 
actual test points superimposed. Three additional starts were made which are 
not documented in the figures. The total cumulative running time was 96 min. 

Prior to rig rotation, helium pressures ranging from 662 to 1827 kPa absolute 
(96 Co 265 psia) were applied to check the integrity of the seal installations 
and prov' de zero speed data. 

During operation, steady-state data were sequenCially taken at 3665 rad/a 
(35,000 r/min), 4188 rad/s (40,000 r/min), 4712 rad/s (45,000 r/min), and 
5235 rad/s (50,000 r/min) with helium supply pressures ranging from approxi- 
mately 552 to 1172 kPa absolute (80 Co 170 psia). Drain pressures downstream 
of both seals were maintained between 448 Co 552 kPa absolute (65 to 80 psia). 
The helium supply temperature varied between 38 and 43°C (100 and 110"^?). 

As indicated on Figure 3-3, virtually all of the test points were in Che high 
friction region of the operating map. At lower speeds of 3665 rad/s 
(35,000 r/min) and 4188 rad/s (45,000 r/min), this was necessary to maintain 
a positive pressure drop across Che seals. At Che higher speeds of 4712 Co 



■''Figures are presenCed consecucively, beginning on page 3-48. 



3-3 



')/■)') ra«l/'i (45,000 l.o 55,000 r/mui), [iro[)iirt lonal 1 y hij'.luT prcisurci wcri' 
maintained in an (^tfort to add dampinj;, to the rotur fiyjtnm to ('timbat the 
vibrat^ion problom. 

r)(^'ipir,(> ihono a* t.(>mpt. <j , vit)rat.i(jn!i w(!re cncouni crod at approximaf I'l" 
5235 rad/'j (50,000 r/min) and bocariK? worr.e. ,in 'jpofid inc. re.ifiod to 5/5'J rati/'i 
(55,000 r/min). While aeveral data points were taken at the latter lipeed, the 
tester was tripped out when the outboard ueal temperature abruptly increased 
signityinp, a rub. l-'iRure 3-4 is a strip chart recording', of the temperature 
excursion. A torquti check revealed that the shaft was ti|^,ht. (;onf!e']uent 1 y , 
no additional running, was attempted. An analysis ot the tailed hardware is 
f^iven in Section 3.5.1. 

In general, the facility and instrumentation worked well; howev(?r, sev(>ral 
problems did develop. The two capacitance probes obsi'v ap, the s(ial runner 
rubbed briefly and shorted out due to rotor vibration and a greater than 
anticipated loss of gap during initial chill-dQwn= The failures occurred 
shortly a.ter the first start. Consequently, no usable probe data was logged. 
Secondly, the Ventu i. flowmeter, set to measure the leakage from the outboard 
seal was sized anticipating larger flows and resulted in pressure drops t jo 
low to measure. Both problems were corrected before the second seal test. 

Figure 3-5 shows both measured and predicted flow as a function of pressure 
drop across a single seal ring. The data is for the 7.ero speed case with the 
tester chilled down. The theoretical predictions are base ^ on the latest 
model incorporating both viscous and inertial pressure drops. (See Appen- 
dix A.) The four theoretical curves cover the range of anticipated radial 
clearances and were calculated using the data giv'en in the block insert in the 
figure. Because of the flowmeter problem, the flow through the individual 
sea! could not be separated. The experimental flow plotted equals one-halt 
the total measured supply flow. The data in the block insert also applies to 
the experimental flow curve except for the eccentricity ratio and discharge 
coetricient which could not be measured. 



3-4 



Because the film thickness was not measured, no correlation between measured 
flow and clearance was possible. However, cor elating the measured flow with 
the theoretical flows showed that the experimental data follows Che predicted 
curve Jor f radial film thickness of 0.020 mm (0.0008 in.) at low flows and 
C.023 mm (0.0009 in.) at higher flow conditions. Given the room temperature 
clearances of 0.015 mm (0.0006 in.) for the outboard seal and 0.008 mm 
(0.0003 in.) for the inboard seal, the measured flow seems reasonable. 

Figure 3-6 shows the surface temperaturr of the seal rings during the same 
test. The temperatures increase between 65 and 85°C and (117 and 153°F) from 
low flow to high flow conditions. This demonstrates the strong warming effect 
that the helium flow has on the seal rings. It is somewhat surprising that the 
inboard ring registers slightly higher temperatures than the outboard ring. 

Figure 3-7 contains a series of experimental flow curves at different tester 
speeds. Again, these represent one half the total measured flow. Theoretical 
curves are also depicted covering four seal clearances. The maximum flow rate 
occurred at 5759 rad/s (55,000 r/min) at a pressure drop of approximately 
650 kPi (94 psi) and is equal to 0.0017 kg/s (21 scfm). 

The experimental curves show a definite decrease in flow as speed is 
increased. This is probably due to a decrease in clearance caused by centri- 
fuge, and thermal growth of the runner. The anticipated decrease in clearance 
due to centrifugal runner growth is about 0.001 mm (0.00004 in.) for each 
speed increment of 524 rad/s (500C r/min). Thus, if the flow at 3665 rad/s 
(35,000 r/mj.n) corresponds to a clearance of 0.026 mm (0.00106 in.), the 
curve at 5235 rad/s (50,000 r/min) should match up with a curve at 0.023 mm 
(0.0009 in.). The actual data at 5235 rad/s (50,000 r/min) follows a clear- 
ance line of 0.021 mm (0.00083 in.), indicating a slightly greater than 
expected clearance decrease. Also, the data at 5759 rad/s (55,000 r/min) does 
not follow the trend at all. 

Figures 3-8 and 3-9 present the corresponding seal ring temperatures for the 
inboard and outboard seal rings respectively. Again, there is a very signif- 
icant warming effect as flow increases at higher pressures. It is also readi- 
ly apparent that the rate of change of temperature with pressure is much 



3-5 



higher at lower flows and tends to Level off at higher rates. At constant 
pressure, both figures tend to show slight decreases in lemperature as speed 
is incr«>ased. This goes along with the decreased clearance, and thus 
decreased flow, due to centrifugal and thermal growth of the runner. The data 
at 5759 rad/s ('J5,000 r/min) deviates from the trend. 

Figure 3-10 compares inboard and outboard seal temperatures at 4188 rad/s 
(40,000 r/min) and shows that the inboard t«nds to be about 5 to 10°C (9 to 
18°F) colder than the outboard at constant pressure. This is also typical of 
the data at other speeds. 

Since neither the individual seal flow rates nor the actual operating clear- 
ances were measured for seal set No. 1, it was difficult to make any more 
definitive correlation. 

3.2.2 Seal Se t No. 2 

Figures 3-11 through 3-16 give a time history covering the three days over 
which the seals were tested. The curves depict tester speed and helium supply 
pressure. Figure 3-17 provides an operating map covering the entire test. As 
mentioned in Section 2.0, the seal principally operated in the high friction 
range. 

Steady-state data were accrued over five runs resulting in data at 3665 rad/s 
(35,000 r/min), 4188 rad/s (40,000 i/min), 4712 rad/s (45,000 r/min), and 
5026 rad/s (48,000 r/min). Speed was limited to 5026 rad/s (48,000 r/min) 
due to sharply increasing rotor vibrations. Seal set No. 2 completed all 
subjected tests without evidence of failure of any kind. 

The first four runs covered a range of supply pressures from 586 to 1069 kPa 
absolute (85 to 155 psia). The fifth run conducted after a series of high 
acceleration rate runs included additional data points at 4712 rad/s 
(45,000 r/min) and reached a helium supply pressure of 1482 kPa absolute 
(215 psia) . 



3-6 



In addition to varying Che helium supply pressure aC each speed, the outboard 
seal drain pressure was also varied including values of 517, 310, and 103 kPa 
absolute (75, 45, and 15 psia). The inboard drain j -ire was maintained 
between 517 and 551 k.Pa absolute (75 and 80 psia). The upst eam heliun supply 
temperature was between 38 and 43°C (100 and 110°F). 

A zero speed run was also made varying helium supply pressures from 648 to 
1820 kPa absolute (94 to 264 psia). Both seal drain pressures were maintained 
at approximately 517 kPa absolute (75 psia) during this particular test. 

Not counting the high acceleration rate runs which "-e discussed in 
Section 3.5, a total of eight runs were made resulting in a total accumulated 
test time of 261 min. No problems were encountered and the seals were removed 
intact and undamaged. Also, no problems developed in the instrumentation 
which yielded good supply and drain flow data. While the capacitance probes 
functioned properly, later analyses revealed potential inaccuracies in Che 
readings due to thermal distortions of the seal housing. 

Zero speed flows are given for Che inboard and outboard seals in Figures 3-18 
and 3-19, respec*- ' vely. Both predicted and measured data are given. The 
Cotal flow through the seals is approximately 0.0068 kg/s (85 scfm) at a pres- 
sure drop of 1250 to 1300 kPa (181 to 189 psi). This is slightly less than 
the 0.0077 kg/s (97 scfm) passed by seal set No. I at the same pressure. 
However, seal set No. 2 ha' slightly lower clearances (see Table 3-1). 

The flow from Che inboard seal dominates by a factor of two to one and closely 
follows the predicted flow curve corresponding to radial clearances of 
0.024 to 0.027 mm (0.0009 Co O.OOll in.). This represents an increase in 
clearance of 0.018 mm (0.0007 in.) from ambient to test conditions due to 
conCraction of the runner from LN2 at the inboard end. The outboard seal flow 
runs along the 0.020 mm (0.0008 in.) predicted flow curve at lower pressures 
and falls to 0.017 mm (0.0007 in.) at higher pressures. It shows a more 
modest clearance increase of 0.007 to 0.010 mm (0.0003 to 0.0004 in.) from 
ambient to test conditions. 



3-7 



Figure 3-20 shows the seal surface temperatures for the zero speed test. As 
in the previous test, a substantial temperature increase is registered as the 
pressure drop is increased. Although of a slightly lower magnitude, 40 to 
SO'C (72 to gO^F), as opposed to the 65 to 85°C (117 to 153°F) in seal set 
No. 1, the inboard seal again shows the higher temperature. The reason in 
this case is the larger flow passed by the inboard seal. This may also be the 
case with seal set No. 1. 

The flow curves for the inboard real at the different test speeds are given in 
Figure 3-21, The first four runs comprise the data at 3665, 4186, 5026 rad/s 
(35,000, 40,00 , 48,000 r/min), and the lower pressure data at 4712 rad/s 
(45,000 r/min). These runs produced the maximum flow rate of 0.0018 kg/s 
(23 scfm) at a pressure drop of approximately '+30 kPa (62 psi) and show the 
expected trend of decreasing clearance as speed increases. Again, the magni- 
tude of the decrease is more than can be explained by centrifugal growth 
alone. The actual decrease between 3665 and 5026 rad/s (35,000 and 
48,000 r/min) appears to be approximately 0.010 mm (0. ■!04 in.) while the 
predicted amount is 0.0025 mm (0.0001 in.). Clearly, additional effects 
(thermal growths) are taking place. 

A fifth run was made at a speed cf 4712 rad/s (45,000 r/min) during which a 
pressure drop of 940 kPa (136 psi) WdS applied to the seals. Given the 
prevailing drain pressure of 544 kPa absolute (79 psia), the absolute pres- 
sure slightly exceeded the design point of 1379 kPa absolute (200 psia). The 
resulting flows are much lower and bear little correspondence with the first 
four runs. 

Figure 3-22 presents the inboard seal temperature variations corresponding to 
the flow curves in Figure 3-21. A high temperature gradient with pressure 
drop exists at low presoures that gradually decrease j pressure increases. 
However, no clear-cut trend ex sts among the first four runs at different 
speeds. The high-pressure run - 4712 rid/s (45,000 r/min) shows quite a bit 
lower temperature than the earlier run at the same speed. 

Outboard seal leakage flows art; plotted in Figures 3-23, 3-24, and 2-5, Four 
speeds are represented in sach figu'-e with two runs at different pressure 



3-8 



ranges at 4712 rad/s (45,000 r/min). Figure 3-23 corresponds Co an outboard 
drain pressure of 517 kPa absolute (75 psia). Figure 3-24 correaponds to a 
drain pressure of 310 kPa absolute (45 psia) and Figure 2-5 to a drain pres- 
sure of 103 kPa absolute (15 psia). The results here are quite different than 
those shown for the inboard seal and yield the following observat iong : 

'.. The flows are lower than those of the inboard seal indicating smaller 
clearances . 

2. The flows increase only slightly with pressure. This indicates that 
the clearances are closing as the supply pressure increases. 

3. There is no discernable trend among flow lines at different speeds. 
The loss of clearance due to centrifugal growth is most likely over- 
shadowed by the flow-related temperature effects. 

Outboard seal temperature plots are given in Figures 3-25, 3-26, and 3-27 
corresponding to the data given in the flow plots. Previously astablished 
trends continue to prevail with higher temperatures, lower gradients at higher 
pressure drops, and the lack of a definite trend as speed is increased. 

1. 

Figure 3-28 is a plot of the outboard seal temperature and flow during the 
"5 first run at 41887 rad/s (40,000 r/min) at each of the three drain pressures 
Jii» versus pressure drop across the seal. Figure 3-29 plots the same temperature 
^ versus the supply pressure. Referring to Figure 3-28, the flow follows a 
J^ straight line as a function ot pressure drop across the seal and appears to be 

independent of drain pressure. The corresponding seal temperatures, on the 
I other hand, all cover about the same range of -40 to -25°C (-40 to -15°F) and 

seem to be independent of the increased presssure drop permitted by the incre- 
I mented decreases in drain pressure. Figure 3-29 shows a different relation- 

ship. The flow curves appear as parallel lines as p.xpected. The temperature 

L curves are much more closely grouped together and, in fact, coincide at -34°C 
(-30°F) indicating that the outboard seal teinperatures have a stronger depend- 
ency on the total seal supply pressure than the pressure drop across the 
i 

I* outboard seal. This tfnds to indicate a strong thermal interaction between 
the outboard seal and the inboard seal with the larger, more dominant flow. 



i 

L 



3-9 



Figures 3-30, 3-31, and 3-32 are plots of seal film thickness versus time for 
the five test runs being analyzed. As mentioned previously, evidence of 
substantial thermal distortion of the seal housing was observed during the 
chilL-down process which appears to have caused inaccuracies in the absolute 
measurement of the film thickness. This is particularly evident in the verti- 
cal outboard seal curve which shows nearly three times the expected film 
thickness. 

3.2.3 Seal Set No. 3 

Figures 3-33 and 3-34 give time history plots of supply pressure and test rig 
speed, while Figure 3-35 gives an operating map. Principal data was acquired 
during three runs taken over a 240-min period. Two additional starts were 
made. The total cumulative running time was 236 min. 

Steady-state data were taken at 3665, 4188, and 4712 rad/s (35,000, 40,000, 
and 45,000 r/min). Attempts were again made to reach higher speeds but 
vibrations were encountered at approximately 5026 rad/s (48,000 r/min). It 
was decided to limit operation to 4712 rad/s (45,000 r/min). 

A broad range of supply pressures were achieved from 654 to 1482 kPa absolute 
(95 to 215 psia). Outboard seal drain pressures were incremented at each 
speed covering values of 517, 310, and 103 kPa absolute (75, 45, and 15 psia). 
No zero speed data was taken with this seal set. 

The steady-state tests were completed without problems and were followed by 
high acceleration rate tests. After 40 fast starts and during a short 
steady-state run to check operation, a rub was encountered on the outboard 
seal. Figure 3-36 shows the strip chart data documenting the increase in seal 
ring temperature and the subsequent shutdown. Two attempts to restart 
resulted in immediate increases in temperature indicating that significant 
damage had taken place during the first occurrence. The rig was then torn 
down and inspected. A fair amount of contamination had entered the supply 
cavity and probably initiated the rub. Subsection 3.6 provides additional 
details of the inspection and an analysis of the failed parts. 



3-10 



Figure 3-37 provides Che inboard seal flow data at the three test speeds; 
predicted flows are superimposed. The maximum flow is approximately 
0.0026 kg/s (33 scfm) and occurred at 3665 rad/s (35,000 r/min) at a pressure 
drop of 630 kPa (92 psi). The measured flows show a decrease with speed, 
although the amount is more than expected based on centrifugal growth of the 
runner. 

Figure 3-38 shows the corresponding inboard seal temperatures. The data shows 
a definite decrease in temperature as speed increases and presents a clear 
trend unlike the corresponding plot (Figure 3-22) from seal set No. 2. 

Flow data from the outboard seal at different drain pressures are shown in 
Figures 3-39, 3-40, and 3-41. The flows range between 0.0008 and 0.0012 kg/s 
(10 and 15 scfm) going from low to high pressure drops. Speed appears to have 
no appreciable effect. The small variation with pressure seems to indicate a 
decrease in operating clearance as observed in previous tests. 

The temperature plots for the outboard seal are given in Figures 3-42, 3-43, 
and 3-44. As with the inboard seal, the temperatures are generally lower at 
higher speeds. 

3.2.4 Seal Set No. 4 

This test was comparatively short due to a rub on tHe inboard seal. One start 
was made and the cumulative test time was about 20 min. Figure 3-45 is an 
operating map while Figures 3-46 and 3-47 present time history plots showing 
seal supply pressures and tester speeds. The time history plots are expanded 
to include about 2 hours of data prior to and after the running portion of the 
test. 

The strategy for this test was somewhat different than for the fivsu three 
tests because of the test rig modifications. The most important was the laby- 
rinth seal. With this in place, it was hoped that the vibration problem would 
be solved and ' : tester would be able to run at the full speed of 7329 rad/s 
(70,000 r/min). Because of the importance of this achievement, testing at 



3-11 



multipLe pressures at low apeeds that had been done in previous tests was 
postponed. 

The test run consisted of an initial ascent to 3141 rad/s (30,000 r/min) 
followed by incremental increases of 524 rad/s (5,000 r/min) to about 
5759 rad/s (55,000 r/min). The supply pressure was held constant at 827 kPa 
absolute (120 psia). The relatively low value was chosen to minimize the risk 
of friction lock-up. The inboard seal drain pressure wau held at 
approximately 637 kPa absolute (92 psia) while the outboard seal drain pres- 
sure went from 483 to 310 kPa absolute (70 to 45 psia) as the test proceeded. 

Operation at 5235 and 5759 rad/s (50,000 and 55,000 r/min) showed that the 
vibrations were lower than in previous tests. However, the outboard seal 
clearances which were being measured using the embedded probes also revealed 
low film thicknesses of approximately 0.008 mm (0.0003 in.) and a slight 
vertical eccentricity reducing the distance between the seal ring and the 
runner. To provide the best chance for the seal to center itself, the supply 
pressure was dropped slightly to 724 kPa absolute (105 psia) prior to increas- 
ing speed above 5759 rad/s (55,000 r/min). Speed was increased to 5968 rad/s 
(57,000 r/min). Shortly after, a rub occurred on the noninstrumented inboard 
seal ring and the tester was tripped out. Figure 3-48 shows a strip chart of 
the failure. A torque check on the unit revealed that damage had taken place 
and the test was ended. 

Despite the narrow span of operating conditions and the seal failure , a fair 
amount of good data was recorded including a running measurement of seal film 
thickness and eccentricity from the embedded capacitance probes. The film 
thickness measurements were very important because they provided one of the 
crucial links between the theoretical model and the actual seal performance. 
Previous tests fell short in achieving the necessary accuracy in film thick- 
ness measurements. 

To maximize the data yielded by the test, the analysis was expanded to include 
not only the actual run which essentially consisted of a very slow speed sweep 
at constant supply pressure, but also a pretest and post-test period each 
consisting of about 2 hours. Full data scans including measurements of pres- 



3-12 



\m 



fiurea, flows, temperatures, and film t hicknesijeu were taken diirin^', both inter- 
vals. The pretest period showed the seals in the firi.il stages of chill-down, 
prior to introducing a significant flow of helium and the subsequent effects 
on ring temperature and clearances as Che pressures were increased to the 
levels applied during the run. While a high-pressure zero speed run was not 
made, as had been done for seal sets Nos. 1 and 2, sufficient data points were 
covered Co establish a zero speed flow-pressure curve. This also provided a 
zero speed datum at the run pressure to help characterize the effects of speed 
on film thickness. The post-test data is a little less revealing because the 
helium was shut off shortly after the test rig was tripped out. However, 
several data points were taken before this occurred. These consisted of a 
repeat of several flow-pressure points which permitted a comparison with 
prerun data. The results clearly demonstrated the increase in clearance of the 
inboard seal caused by the rub. 

Because the test consisted essentially of one run rather than a number of 
runs, as was the case in previous tests, the data is most easily reviewed and 
is hence presented in a chronological or time history frormat. Specific plots 
of one variable against another are presented as required to illustrate key 
parametric relationships. The two independent variables a^e speed and pres- 
sure. Time history plots of these were presented in Figures 3-46 and 3-47. 
Because the seal drain pressures were different than the previous tests and 
the outboard drain pressure tended to vary somewhat, differential pressures 
across both seal rings are plotteu along with the absolute supply pressure. 
Leakage flows are plotted in Figures 3-48 and 3-49. 

During the initial part of the pretest period (-100 to -60 min) the supply 
pressure was held at the relatively low value of 310 kPa absolute (45 psia) to 
maintain a small but positive flow th'-ough both seal rings while the tester 
was undergoing final chill-down. Subsequent to this, the seal supply pressure 
was increased to 827 kPa absolute (120 psi?) in preparation for starting. 
During the entire period, the outboard seal flow was quite a bit larger than 
the inboard flow. 

Figure 2-6 presents a plot of the measured pretest zero speed flows versus the 
predicted flows at various clearances. Most of the data from the outboard 



3-13 



fisal toLlowg the 0.0305 mm (0.0012 in.) clearance line while chat from thH 
inboard seal tended toward even higher indicated clearances. Both meajjurod 
flows are higher than the corresponding data from seal sot No. 2 (see 
Figures 3-18 and 3-19). A direct comparison, however, is difficult because of 
the various differences between the seals and instrumentation. One of the big 
differences which supports the existing data is that seal set No. 4 had room 
temperature clearances which were nearly twice those of seal set No. 2. A 
difference which drives the argument in the opposite direction is Che differ- 
ence in boundary temperatures and their effect on the seals and runners. 

Seal set No. 4 and runners were relatively warm because of Che labyrinth seal 
(see Figure 2-18) which prevented LN2 from flowing into the inboard seal drain 
cavity. Seal temperatures were in the -10 to +10°C (14 to 50°F) range. Seal 
set No. 2 was much colder because in that installation Che inboard end of the 
runner was in direct contact with Che LN2 flow (see Figure 2-15). For Che 
same pressure range, their temperaCure varied becween about -55 and -30°C (-67 
and -22 F). Thus, because seal set No. 2 and the riinner were colder, their 
increase in clearan:e from room temperature Co the CesC condicion would be 
greaCer. 

During the actual test run the outboard seal differential pressure rose 
initially from approximately 350 Co 450 kPa (51 Co 68 psi) due Co decreases in 
Che drain pressure. It later dropped back to about 400 kPa (58 psi) when Che 
supply pressure was lowered. The inboard seal pressure drop stayed constant 
aC 200 kPa (29 psi) unCil the supply was lowered when ic dropped Co abouC 
90 kPa (13 psi). 

Boch flows dropped signif icanCly during Che run as speed increased. The 
oucboard seal went from abouC 0.0013 Co 0.0005 kg/s (16 to 6 scfm). The 
inboard seal went from 0.0006 kg/s (8 scfm) to zero flow which coincides with 
the occurrence of the rub and indicates a complete loss of clearance in the 
seal . 

While there is some variation of flow with the changes in outboard seal pres- 
sure drop, and Chere may be some effect on both seals due to temperaCure 
changes, Che principal cause of Che changing flows is Che decrease in clear- 



3-14 



I 
I 
I 

E 
I 
[ 
C 
I. 

L 
L 
I 
I 
I 
i 
L 
L 
I 
L 



ance as speed increased, figure 2-7 prd.eiits a ploc ot tlow versus pressiirt* 
drop tor Che test run illuacrating the ettect;. 

Both seal flows increased after the tester was shut down. The flow through 
the outboard seal returned to >^802 "if the level recorded prior to the run at 
the same differential pressure. The flow througli the inboard seal, however, 
went up dramatically reaching twice the level achieved before operation. The 
net increase through the inboard seal clearly reflects the increase in clear- 
ance due to the material worn away during the rub. 

Refocusing on the flow curves for the outboard seal in Figures 2-6 and 2-7, 
another anomaly is observed. The actual measured (average) film thicknesses 
are noted for many of the data points and are considerably smaller than the 
radial clearances corresponding to the predicted flows. The ratio of 
measured-to-predicted seal clearance is approximately 0.64 with the theory 
tending to overstate Che clearance required to achieve a given flow. A possi- 
ble explanation is that the "extra" flow is bypassing the normal path and 
leaking out between Che axial sealing surfaces. Unfortunately, the data does 
not allow a more definitive explanation. 

Figure 3-50 plots Che surface CemperaCures of both seal rings. While Che 
outboard seal is slightly warmer, both seals follow a parallel path with a 
difference of about 7 Co 8°C (13 Co 14°F). Boch are observed Co decrease in 
Che early porcion of Che pretest period due to steadily declining temperatures 
in the test rig. After Che helium flow is increased (-60 min Co start-up 
time), the temperature incieasea and attained values slightly greater Chan 
Chey had aC Che beginning of Che precesC period. This resulCed from Che warm- 
ing effect of the helium flow. Previous discussions pointed out the substan- 
tial differences between the temperatures of these seals and Chose of seal set 
No. 2 (also seal sets Nos. 1 and 3) with Che higher CemperaCures being 
achieved by beCter isolation of the seals and runner from the cryogenic bear- 
ing fluid. 

Both the gradual chill-down and Che subsequenC warming effecC can be observed 
in Che behavior of the outboard seal film thickness during Che pretest period. 
This is shown in Figure 3-51. The corresponding eccentricities are shown in 



3-15 



Fi^'uro 3-52. Initially, when r.h(> soal t ('mpcrature way ducroasiiu' the avcraj'c 
tilm Chicknesa ineroaaed alightly. Inveriio.ly, during t.lio later prnttMU period 
(-60 min and starL-up) che eloarance closed up again a*; t lie t cmpcraturf 
increased. 

The change in seal clearance ia directly due t;o t;he tGmperatur(! chan>»e in t,he 
seal ring and runner. If the temperature charges in both parts are the same, 
the rate of change of the radial seal clearance is about 2.1 x 10 mm/°C 
(4.6 X 10" in./°F). This is based on expansion coefficients of 12.6 x 
10"^ m/m-''C (7.0 x lO"^ in./in.-°F) and 4.3 x 10"^ m/m-°C (2.4 x 
10" in./in-^F) for Inconel 718 and I'-5N carbon graphite respect; ively. 
Ignoring the discontinuity at -50 min, the outboard seal temperature changes 
about 14°C (57.2°F) during the pretest period. Assuming equal changes in the 
runner temperature, the outboard seal clearance should have changed about; 
0.0029 mm (0.00011 in.). This compares well with the measured change which is 
•^0.004 mm. 

Referring again to Figures 3-48 and 3-49, both seal temperature and clearance 
decrease substantially as speed is increased during the test run. The princi- 
pal effect is the centrifugal growth of the seal runner which causes th(> seal 
clearance to decrease accordingly. This, in turn, causes the leakage flow to 
go down and, hence, the seal ring te.nperatures e,o down. 

Figure 3-53 plots the measured average film thicknesses of the outboard seal 
as a function of speed. Also plotted is the predicted film thickness based on 
a zero speed film thickness equal to 0.016 mm (0.0006 in.) and a full speed 
7329 rad/s (70,000 r/min) radial growth of 0.011 mm (0.00045 in.). Again, 
the data shows very strong correlation between measured and predicted values. 

While the decrease in temperature is due to the decrease in flow, the effect 
is not as easy to substantiate by calculation. However, a comparison of data 
points during the test run with data points recorded during the pretest period 
which sf.ow equal flows, also reveal data points with nearly equal temper- 
atures. For example, the flow through the outboard seal at 5 min after 
start-up is about 0.001 kg/s (8 scfm). This is the same as the flow at 



3-16 



'♦() min. Tho outboard <um\ ffimpcrat are i ■; approximately ')"',' (4l"l'') at both 
t imar,. 

What doesj not appear, or at loaat i;] not obvious, ifj the nttuct ot t ht; temper- 
ature decrease during, the run on the tilm thicknofj'i. The •jimpic calrularion 
contained in the previous diseussioni would predict aa inLrea'iu of approxi- 
mateLy 0.003 mm (0.00013 in,) baaed on the measured temperature drop of 10"c 
(18°?). However, since this change is about one third of the change dui.' to 
centrifugal growth, its effect could bo easily overlooked. 

Toward the end of the test run the inboard seal temperature began to increase, 
showing a 4°G (7°F) change over about i min. This correspond'; with the time 
at which the seal dropped to near zero clearance and signifies the start ol 
the seal failure. 

Immediately after the tester was shut down, '.oth oe"' temperatures and the 
outboard seal film thickness returned to virtually the same values as they had 
immediately before the run. After the helium supply was shut off at about 
30 min, the tempe.-atures dropped nearly 70°C (126°F) as the seal components 
were chilled by the cryogenic section of the test rig. This, in turn, 
resulted in a 0.016 mm (0.0006 in.) increase in film thickness which is the 
expected amount based on previous rates. 

3.3 Dynamic Behavior Durin g S eal Testing 

While the tests were not structured to examine the dynamic behavior of the 
seals, it is an aspect of considerable importance. Dynamic data was thus 
monitored, recorded, and later analyzed. Basic characteristics such as the 
size of the runner orbit and whether the seal rings remain locked or tended to 
whirl were disclosed. 

For the first three seal tests, the instrumentation consisted of six capaci- 
tance probes observing the seals and the runner. During the first test, no 
data were gathered because two of the probes failed and several others were 
forced out of their usable operating range. The problems were coriected and, 
despite thermally induced offsets which caused bothersome errors in static 



3-17 



jM)3icion mea!jur(>m(>nf !i, the 'jccond and third 'U>al Leot-i pruvcd morn sih i^i'-iiil^ul 
and yielded good dynamic data. The tOurch ueal tost al-io resu,ll,t'd in ^'/jud 
dynamic miormat: i on t rom the outboard 'itsal containing?, the tour (imbt'dded pro()f"j 
and the two probes retained to ob.'iorve the runner. No mea.'juroment u wore made 
on the inboard seal ring. 

3.3.1^^^ Seal Set No. 2 

With only a tew exceptions which will be noted, the seal rings were motionless 
during steady-state operation. The runner orbit varied in size between 0,004 
and 0.010 mm (0.0001'j and 0.0004 in.) reaching the larger sixe at higher 
speeds. The orbit consisted only of the synchronous and its harmonic frequen- 
cies. 

[•'igure 3-54 shows one of the c.« >es where seal motion was detected. This 
occurred at 3665 rad/s (35,000 r/min) when the outboard .eal drain pressure 
was lowered to 103 kPa absolute (15 psia). The supply pressure was 655 kPa 
absolute (95 psia). Both seals show approximately 0.013 mm (0.0003 in.) of 
straight-line motion, the inboard seal moving vertically whilt the outboard 
seal vibrated horizontally. The seal amplitudes grew worse with supply pres- 
sure and the condition shown was the highest su[:;;ly pressure attempted. 

While the seal probe settings seemed to checR out, the unu^ju.jl nature of the 
indicated motion (i.e., absolutely straight-line) cast i somr doubt on thti 
validity of the data. 

Figure 3-55 shows seal and runner orbits at 4188 rad/s (40,000 r/inin) with a 
supply pressure of 862 kPa absolute (125 psia) and equal drain pressures of 
517 kPa absolute (75 psia). This data point was taken immediately following 
the one illustrated in Figure 3~52. A small amount of th^ straight-line seal 
motion remains. The seals were motionless at all other flow conditions at 
this speed. 

Figure 3-56 depicts the seal runner orbits during the first data point of run 
No. 3. Speed was 4712 rad/s (45,000 r/min), supply pressure 938 kPa absolute 
(136 psia), and drain pressure 517 kPa absolute (75 psia). Some seal motion 



3-18 



is observed. At siirceeding data points, the rings again locked and remained 
motionless . 

3.3.2 S eal Set No. 3 

The rurrer from the No. 2 test was reused and resulted in similar runner 
orbits. Tne seal ring motion, however, was quite different. While the 
outboard seal was motionless, the inboard seal exhibited an orbit with ampli- 
tudes at times reaching 0.030 mm (0.0012 in.). 

Figures 3-57, 3-58, 3-59i, ai.d 3-60 show four data points at 3665 rad/s 
(35,000 r/min). The scales in the figures are the same as those in 
Figure 3-54. The first two represent supply pressures of 655 and 1206 kPa 
absolute (95 and 175 psia), respectively. Both drain pressures were 517 kPa 
absolute (75 psia). The second of these shows a narrow elliptical orbit on 
the inboard seal which grew in a uniform manner as the pressure was increased. 

Figures 3-59 and 3-60 represent the data runs where the outboard drain pres- 
sure was reduced to 310 and 103 kPa absolute (45 and 15 psia), respectively, 
and the supply pressure was raised to the maximum value of 1206 kPa absolute 
(175 psia). In each case the inboard seal orbit progressed from small ampli- 
tudes to the sizes shown as the supply pressure was increased. 

Figures 3-61 and 3-62 show similar occurences at 4188 and 4712 rad/s 
(40,000 and 45,000 r/min). In these cases, amplitudes at low supply pressure 
grow from about one-half the amount shown in the photo, to full size. In these 
cases the amplitudes are very large and could easily have resulted in a seal 
ruD. They are clearly unacceptable. 

While the mechanism causing the whirl is unknown, the following summarizes its 
salient aspects : 

1. The whirl is synchronous although some harmonic activity is present. 
It must therefore be a result of runner excitation. 



3-19 



2. The seal amplitude is much Larger than Che runner motion inciicatinj» 
an amplification and a system resonance having to do with the seal 
ring may be involved. 

3. The whirl is obviously very sensitive to the surrounding fluid condi- 
tions, C'g-r supply and drain pressures and also to speed since the 
amplitudes become much worse when speed increases from 3665 to 
4188 rad/s (35,000 to 40,000 r/min). 

4. The whirl did not occur on the outboard ring operating under virtual- 
ly identical conditions, nor did it occur on either of seal set No. 2 
rings. This suggests that there is something unique about the seal 
set No. 3 inboard ring. 

5. The operating map showed that the seal operates in the high friction 
region and the frictional forces (based on a friction coefficient of 
0.20) should exceed the hydrodynimic forces and prevent Che seal from 
moving. There was concern that problems may develop due to the 
inability of the seal rings to move. Clearly this was not the case 
for the seal in question. Its actual behavior suggests that the 
frictional forces are not being developed and that the seal may be 
lifting off enough, at least, to result in the development of a 
partial lubricating film under the radial sealing land to lower the 
prevailing friction coefficient. 

3.3. 3 Seal Set No. 4 

No unusual motions w^ire noted during the test. However, with the exception of 
speed, the range jf operating conditions was v=;ry narrow. Supply pressure was 
held constant at 827 kPa absolute (120 psia). Hence, the conditions under 
which seal set No. 3 showed, large whirl amplitudes were not encountered. 
Speeds covered the range from 3141 to 5968 rad/s (3C,000 and 57,000 r/min). 

Figures 3-63 through 3-67 each show displays of the runner orbit and the 
embedded probe output superimposed on the same photo. Speeds range between 
3665 and 5759 rad/s (35,000 and 55,000 r/min). The runner display shows a 



3-20 



normal orbit using vertical and horizontal probes. Over the <5peed range it 
indicates a modest growth from an orbital diameter of approximately 0.004 to 
0.008 mm (0.00015 to 0.0003 in.) 

The disy'^ys resulting from the embedded probes do not show seal cirbits. Each 
one was generated by a pair of embedded probes. The probes are not 90° to each 
other, rather, they are 150° apart. Thus, one display represents the vertical 
components of the ring motion while the other shows the horizontal motion. 
The arrangement is illustrated in Figure J-68. The display format provides 
several pieces of information. Measurements 'a' and 'b" represent the average 
film thickness at each probe; 'c' is the arithmetic average of 'a' and 'b', 
and represents the average radial film thickness in the direction of the 
probes; 'd' is half the difference between 'a' and 'b', and represents the 
eccentricity in the direction of the probes. The x- and y-axes of the scope 
display represent the zero gap values and the point at which the seal rubs the 
runner. They are shown by the heavy white lines in the photo-^- and are the same 
for both pairs of probes in Figures 3-63, 3-64, and 3-65= For Figures 3-66 
and 3-67, the y-axis for the vertical probe display is moved to the right to 
prevent the displays from overlapping. All other aspects are the same. The 
position of the runner orbit on the photo is completely arbitrary and has no 
bearing on the other displays. If the ring remains rigid, a relative orbit 
between seal and runner registers a straight line with a slope of -1. The 
whirl amplitude in the gwen direction is equal t :^ 'e'. 

Referring again to Figures 3-63 through 3-67, it is observed that the horizon- 
tal seal probe display shows as a straight line with x and y components which 
are equal to themselves and to the horizontal diameter ot the runner orbit. 
This indicates I'hat all of the relative motion measured by the embedded probes 
is actually seal runner motion and the seal ring is essentially stationary. 

The vertical seal probe displays are not very straight. This is because ot 
harrronic distortion in the horizontal component of the scope display. 
Figure 3-69 shows the relative seal motion versus time for the four embedded 
probes. The fourth trace shows the distortion. The distorted waveform should 
be equal but out of phase with the third trace which shows the output from the 



3-21 



probe insCalled 180° away. Note the mirror image relaCionship between the top 
two traces representing the other opposed pair of probes. 

Literally interpreted, the f^istortion indicates that the seal ring is flexing; 
however, because there is no evidence of harmonic activity in the other traces 
it is discounted as an instrumentation anomaly. 

Observing the nondistorted component of the vertical display, it can be seen 
that the relative seal motion is again equd to the runner motion, indicating 
that the ring is, in fact, stationary. The n ;i. result is that the outboard 
seal ring stayed motionless at all test conditions. 

3.4 Acceleration Testing 

High acceleration rate tests were performed on seal set Nos. 2 and 3. Each 
test consisted of starting the tester from zero speed <ind accelerating it to 
maximum speed at an average rate of 152 m/s (SOO ft/s ). Due to the dynamic 
problems, the maximum speed was set at 4188 rad/s (40,000 r/min) for seal set 
No. 2 and approximately 4712 rad/s (45,000 r/min) for seal set No. 3. T*" i s 
resulted in an acceleration time of appt'oximatc^ly 0.7 to 0.8 s. Supply and 
drain pressures were set prior to the run. The starts were achieved by open- 
ing the turbine solenoid trip valve, allowing the unit to accelerate and clos- 
ing the valve using an overspeed trip signal. Figure 3-/0 depicts a speed and 
time cufve typical of these tests. 

Because each fast start was followed immediately by a shutdown, there was no 
way to check steady-state operation during the run to assess any potential 
damage. Therefore, after every five fast starts performed in the manner 
described, a slow start was made using manual controls. These runs allowed 
the unit to run at steady-state conditions to permit a cursory performance 
check. 

3.4.1 S eal Set No. 2 

A total of 50 fast starts were performed with average acceleration rates vary- 
ing from appruximately 149 to 198 m/s (490 to 650 ft/s' ). The helium supply 



3-22 



pressure was held aC 931 kPa absolute (135 psia) and both inboard and outboard 
drain pressures maintained at 517 kPa absolate (75 psia). 

The first 30 acceleration runs were done in groups of 5 with slow start stead- 
y-state runs in between. The remaining 20 were performed in two groups of 10, 
again with a periodic steady-state check before and after. 

labia 3-2 gives the test conditions and provides steady-state performance data 
between each group of acceleration runs. While the flow rates do show some 
variation, there is no cbvious sign of deterioration. Also, the post-test 
inspection revealed no wear or other signs of distress. 

3.4.2 Seal Set No. 3 

Forty acceleration runs were conducted on these seals with average rates vary- 
ing between 119 and 171 m/s (390 and 560 ft/s ). Helium supply pressures 
started at 1069 kPa absolute (155 psia) on the early runs and increased to 
1482 kPa absolute (215 psia) on the latter runs. Bot-h drain pressures we- 
held at 517 kPa absolute (75 psia). The acceleration runs were done in groups 
of five. 

The runs proceeded in a normal fashion with no sign of problems through run 
No. 48. Tcjjle 3-3 documents steady-state performance between groups of accel- 
eration runs; no obvious deterioration was noted. A ."^ ^cond slow start, run 
No. 48a, was then maae . The conditions were 1482 kPa (215 psia) and 
4712 rad/s (45,000 r/min). About 15 s into the run, the tester was manually 
tripped when its noise level increased abruptly. The strip charts revi-,'led a 
subsequent increase in the outboard seal ring tempertur unifying a ■: ab (see 
Figure 3-36). Two attempts to restart the tester ^suited in additional 
rubbing witn attendant temperature increases. The test was terminated at that 
point . 

3 .5 P ost-Test Hardware In spect !o ns ^nd Failure Analysis 

Tables 3-1 and 3-4 summarize the condition of the seals and runners before and 
after testing. The former provides key dimensional data while the latter is a 



3-23 



qualitative description. 

An overall evaluation of the post-teat condition of the seal rings and runners 
reveal that they fall into essentially one of two categories: those with 
significant damage and those with no damage at all. 

Three seal rings fall into the first group. All were damaged due to rubs 
between the ring and runner. In all cases the rubs happened suddenly and 
unexpectedly. Fortunately, the rubbing was ver/ obvious in the data and 
permitted quick shutdowns. In fact, in all thrtie cases, the turbine was 
tripped within 5 s of Che start of the rub. While the damage was confined for 
the most part to surface effects, it is obvious by the rate at which it took 
place that the rubs were very destructive and had they been allowed to 
progress for even 10 to 20 s, much more serious damage would have resulted. 

The remaining five seal rings .:onstitute the undamaged group. These went 
through the sam.e battery of tests, as the failed seal rings. Additionally, 
three rings were subjected to a fail 're of tbe adjacent seal ring. None of the 
rings showed signs of wear or other distress on Che mating surfaces with the 
possible exception of some slight polishing over short arcs. 

3.5.1 Seal Set No. 1 

The outboard seal ring rubbed due to high runner vibration at a speed of 
approximately 5759 rad/s (55,000 r/min). Removal of the end cap revealed Che 
outboard se^l drain cavity had a film of black soot on all surfaces. The 
damaged outboard seal ring was tight on the runner due both to wear debris and 
surface damage. It had to be worked loose. Figure 3-71 gives a partial view 
of the seal ring showing heavy rub marks over all surfaces of the bore. Before 
and after bore measurements which are shown in Table 3-1 indicate Chat 
0.023 to 0.028 mm (0.0009 to 0.0011 in.) of carbon graphite had worn away. 
This can be observed in Figure 3-71 by noting that the machined depths of the 
pockets was also approximately 0.023 to 0.026 mm (0.0009 to 0.0011 in.) and 
that the lands had worn down and blended with the bottom of the pockets. The 
seal ring showed no other signs of damage although a slight bit of polishing 
was noted on Che axial sealing land. 



3-24 



The inboard seal ring was Loose on Che runner in its normal position, however, 
was tight when it was pulled over the area damaged by the outboard seal. To 
prevent additiona). damage, LN2 was poured onto the runne"- to decrease its size 
and alleviate the bind. The inboard seal was successfully removed and is 
partially shown in Figure 3-72. While there is some evidence of slight 
polishing over a 11; ^0° arc (shown as the darker areas in Figure 3-72), 
there is no real wear. The few scratches which show in the photo most notably, 
those in the axial direction, are thought to have happened during the disas- 
sembly. Very slight polishing of the axial sealing land was noted. 

Figure 3-73 shows the seal runner illustrating several aspects of the damage. 
First, the rubs are confined to th« middle and right-hand or outboard side of 
the runner. The marks in the middle were caused by rubs with the capacitance 
probes. The marks on the right side are due to the seal. They are heaviest 
towards the edge. These show a continuous pattern alL Che way around and 
correspond with the axial breakdown land on the seal. The marUi inboard of 
that land correspond with Che Rayleigh-step portion cf the sesl . These appear 
as discontinuous skip marks and indicate lighter contact In this area. The 
periodicity is probably attributable to slight high spots built in during the 
grinding of the runner. Numerous longitudinal heaC crtck;) are in evidence in 
all rub areas. Also, several small pieces of Che CungsCen i,arhide coating 
came off in the cracked areas. The latter is illusCraC':.d by the magnified 
view shown in Figure 3-7A (11. 'iX). The other end of the runner is free of rub 
marks . 

3.5.2 Seal Set No. 2 

Both seal rings went through Che CesCing wich no rubbing. The posc-CesC 
disassembly showed no sign of sooC or oCher wear debris. BoCh rings were 
loose and easily disassembled. ExcepC for a slighc bic of polishing over a 
90° band on Che downsCream ef'f^e of Che inboard seal ring and in spocs on Che 
axial sealing land of boch seal rings, neicher ring shows evidence of having 
run at all. Figure 3-75 shows a partial view of the outboard ring. Likewise, 
an inspection of the runner revealed no evidence of any marks. 



3-25 



'^ -h 3 Seal Sjc No. 3 

The uutboard seal ring rubbed unexpectedly while running at steady-state 
conditions at 4712 rad/s (45,000 r/min). 

Removal of the tester end cap and seals ii' >;aled that in addition to umall 
amounts of black soot in both the seal supply chamber and the outboard ;)eal 
drain cavity, thure was several small pieces if leaf mafarialo It was also 
found that one of the two helium supply ports was partially plugged with the 
same leaf material. The failure was thus attribuiirl to the presenci! of the 
contamination. 

Both rings were removed by chl'.ling the runner with LN2 . This was done to 
prevent additional damage. The outboard seal 'ihowed gubsf nni ial Vy less damage 
than the failure of the firut seal. In thi'i case, the wear was concentrated on 
the breakdown land and adjafj.it bearing land on the downstream end of the 
ring. Figure 3-76 shows a typical partial view. While the upstream land 
areas were not henvily worn, a number of heavy scratches were in evidence. 
Figure 3-77 sh' j') the uudan.aged inboara ring. 

As with the twr, previous seal sets, the axial sealing face of seal set No. 3 
rings showed some evidence of slight polishing. Likewise, the mating land 
areas '^n the seal housing showed areas of contact where minute deposit'^ of 
carbon graphite had rubbed off. 

The damage to the runner wa!j less than the first failure., however, it followed 
a similar pattern as shown in Figure 3-78. The rub marks are heaviest in the 
area of the breakdown land and less severe under the adjacent bearing land 
areas. The only other marks under the outboard seal are at extreme upstream 
end and are very slight. A pattern of skip marks appear and predominantly 
longitudinal heat cracks abound. Figure 3-79 shows a magnified view (11. 5X) 
of where several small pieces of the tungsten carbide c ating came off. 



3-26 



i^'ifU 



f: 



3.5.j4 Seal J>et No. 4 

The inboard seal rubbed due to insutficienc radial clearance at 5968 rad/'i 
(57,000 r/min). The uuCboard drain cavity was clean and the outboard seal was 
loose on the runner and easily taken out. The inboard ring was tight to the 
runner and had to be pulled off. 

The damaged inboard ring is shown in Figure 3-80. In addition to signs of 
moderate rubbing across most of the width of the seal and all the way around 
its circumference, a single radial crack extended through a cross section of 
the ring. Thi'j can be seen in the photo between the feed groove and the end of 
the adjacent pocket. 

The outboard seal ring which has the embedded capacitance probes is shown in 
Figure 2-24. It is completely undamaged and shows no sign of having rubbed. 

A photo of the runner is shown in Figure 3-81, Heavy rub marks appear in the 
areas undef the pressure breakdown land and the adjacent bearing land. Next 
to the latter there is some evidence of the "skip marks" noted on the other 
runners and narrow streak-like rub marks under the other bearing land. The 
entire area under the seal ring shows tvidence of polishing. The middle and 
other end of the runner show no marks at all. 

3_. 6 Discu ssio n Of ^h e Res^ul^ t s 

The testing addressed a number of important aspects of the design of the 50-mm 
Rayleigh-step helium biiffer seal. These include: 

• Steady-state operation 

• Fast-start capability 

• Seal life 

• Leakage rates 

• Parametric effects 

- Seal flow path 

- Environmental interaction 

- Supply pressure 



3-27 



- Shaft fipoed 

- Seal clearance 

- Seal temperature 

• Dynamic performance of the geal rings 

• Material considerations. 

3.6 . 1 Steady- Sjtate^O^eraXuin 

The seals proved themselves capable of operating over a wide range of .supply 
pressures and speeds. Supply pressures of up to 1482 kPa absolute (215 psia) 
were applied to three of the four test seal sets at variojij shaft speeds with 
very satisfactory operation resulting. Testing at maximum pressure included 
slow ascents from low to high speeds with maximum pressure applied and appli- 
cation of increasing pressures up to the maximum value at various constant 
speeds. The only indication of a problem which may have been connected with 
high supply pressures, occurred during the testing of seal set No. 3 when the 
inboard seal ring developed a sizable orbit that appeared to get larger as 
'pressure was increased. 

Despite the whirl, the seal ring ran satisfac'"ori ly at 1482 kPa absolute 
(215 psia) at the maximum allowable tester speed. The dynamic consideration 
of this instance is discussed in Subsection 3.4.2. A great deal of running 
was also done at fairly low supply pressures with no evidence of problems. 

Satisfactory seal operation was achieved up to speeds of 5759 rad/ ; 
(55,000 r/min). Operation at higher speeds was precluded because of the 
dynamics problem in the tester which resulteu in large whirl orbits at the 
seal runner. The whirl which had an estimated double amplitude of 0.038 to 
0.051 mm (0.0015 to 0.002 in.) occurred during seal set No. 1 testing and 
caused the failure of the outboard seal ring. Seal set Nos. 2 and 3 were 
arbitrarily limited to lower speeds to avoid repeating the failui-e. 3i\3l set 
No. 4 inboard also failed at about the same sp<?ed as seal set I^o . 1; however, 
its problem occurred for a different reason and under different circumiptancni „ 
The failure was due to a total loss of clearance. This wa -^ well substantiated 
by the data recorded and presented in Subsection 3.3.4. The clearance loss 
was due to two factors which were: 



3-28 



1. Incrna.'linR centri f:ui',al ^',rowrh ot tht' runner 

2. Higher film r.omporatiir(! due to a higher temperature environment. 

To avoid test rig problems, seal set No. 4 and the runner were not expofjed to 
LN2 (see Subsection 3.1). There wa.'i no question that larger machined clear- 
ances would have prevented the seal trom closing up and allowed operation to 
continue. 

In summary, despite the tester-imposed speed Limitation, all indications were 
that the 9eals would have operated succs.'ssf ul ly at the full speed of 
7329 rad/s (70,000 r/min) with up to maximum supply pressure applied. There- 
fore, it is concluded that the design meets the basic speed pressure perfor- 
mance requirements and should receive continued consideration as a viable 
design. 

!• 6.2 Start-Up Perf o rma • ice 

This appears to be one of the lesser demands imposed on the seals. Both seal 
sets that were subjected to the high acceleration rate tests showed no prob- 
lems during or immediately after any of the fast starts. • This is not unex- 
pected given the conditions which apply during a start-up. Helium supply 
pressure is applied to the seals prior to rotation. During the testing, this 
included various pressures up to a maximum of 1482 kPa absolute (215 psia). 
In actual turbopump operation the full design pressure is applied. The pres- 
sure seats the seal rinps and establishes flow, although the rings are proba- 
bly not concentric with the runner. During the initial start, rubbing will 
occur if the seal is in contact with the runner. As the spead increases, a 
hydrodynamic film develops which results in forces tending to center the seal. 
These forces increase as speed goes up until they are sufficient to overcome 
the frictional forces, at which point the seal ring moves to a concentric 
position and the rubbing stops. The required acceleration rates resulted in 
start-up times of approximately 0.77 s from to 4712 rad/s (0 to 
45,000 r/min), the maximum test speed, and 1.20 s for acceleration to 
7329 rad/s (70,000 r/min). Therefore, the time during which rubbing would 
occur is very short. Moreover, because it occurs during the initial part of 
the run, the speeds are lower. This results in less heat and lower temper- 



3-29 



,iture!i at r.he rubbing 'jurtaccfj. All ot thcfje tactors combine ro actually 
tavor the bigh acceleration rate arart'j. 

While no operational problema resulted trom the tast start^j, neither wa;} there 
a buildup of significant wear on the aeal rings. Seal set No. 2 underwent 50 
tast starts while seal set No. 3 was subjected to 40 starts. Kxcluding the 
No. 3 outboard seal which failed due to contamination, the post-test 
inspection (see Subsection 3.6.3) found very little wear on the seal rings. 
In fact, the only real evidence of wear were several short ares in the bore 
area which showed some polishing. This is slightly evident on the left side 
of the photo showing seal set No. 3 ring (Figure 3-77). The tests show that 
the seal design is capable of undergoing multiple high acceleration late 
starts without damage or significant wear. 

3.6.3 Seal La fe 

This readdresses the topics of sceady-state operation and fast start capabili- 
ty but from a slightly different viewpoint. Evaluating seal life pofential 
based on the tests that wert' conducted is a matter of: 

1. Revifv.-ng the life that was achieved in the seals 

2. Evaluating the failures that occurred 

3. Identifying any mechanismh that were present and could have caused a 
fa i lure . 

Table 2-1 presents the cumulative test time and number of fast starts for each 
of the test seals. Seal sets Nos. 2 and 3 were operated for the longest peri- 
odb of time, each one achieving appi oximately 4 h of running. This in itself 
is significant because it represents about 402 of the design life. What is 
more important is that the 4 h logged by each seal set is not really indic- 
ative of their useful life which could have been much longer. T! is is easy to 
see for seal set No. 2. These seals went through the most exte.isive battery 
of tests of the four sets including acceleration testing. At the end of 
tescs^ both the seals and the runner looked as good as before they were test- 
'>d. By all indications they could have been reinstalled and run for an indef- 
inite period. Seal set No. 3 falls into a similar category. It also went 



3-30 



t hr(ni|^,h a liubfit.ant ial 'ichoduln ot tOfU'i which although nor nwitc as Idih',, 
fDutinely achiuved hij^jher supply pri'fjijuroii than scmI r,oA. No. 2. It. operated 
fjuccefifitul, I y right up Co the point at which the outboard 'leal rubbed and t 'le 
teat was terminated, yet the inboard seal waa removed and Like aeal 'jet No. 2 
ahowed almoat no ai^^jna of having run. 

Seal set Non. 1 and 4 logged considerably Leas time than aeal set Noa. 2 
and 3, the former achieving about 1-1/2 h and the latter approximately 20 min. 
Both teats were terminated by faii.irea of one of the rings. However, the 
rings that did not rub ahowod no signs of deterioration and, like seal set 
No. 2, would probably have ru i for a much Longer perioa of time. 

While the three failures that did occur certainly ended the lives of both the 
seals and runners involved, they need to be carefully evaluated to determine 
what they really revealed about the life potential of the seal design. 

As discussed in Subsection 3.6, all three failures share several common char- 
acteristics. The major manifestation of the failures were radial rubs between 
the runner and one of the seal rings. All happenecl very suddenly and unex- 
pectedly. Despite quick shutdowns, significant damage resulted in each case. 
Thus, there is no question that the seal design is very sensitive in this area 
and with existing materials, even the briefest of rubs are to be avoided. 

While damaging rubs were common to all the failures, the triggering mechanism 
was different in each case. Furthermore, each of the mechanisms could have 
been avoided by changes in the design or more careful control of the system in 
which the seals were installed. 

The failures of seal set Nos. I and 3 were caused by outside influences, tne 
former being triggered by runner whirl diameters which were estimated to be 
between 0.038 and 0.051 nnm (0.0015 and 0.002 in.) peak to peak and the latter 
apparently caused by fairly substantial amounts of contamination in the seal 
area. While the exact levels of vibration the seals should be capable of 
handling can be argued, the levels to which seal set No. I was exposed were 
clearly excessive. The amount of contamination was much greater than should 



3-31 



liavfi occurred. Both problomrj c.in be .ivoidijd in tut.urc appl icar i m. Seal si't 
No. 4 failed beeau,'j« ot iiiaut t iciont. c.l (iarii'ico. Thio jI.'JO ir, eauily rc^modicd. 

The only other mechanigma which were observed to have potential lor limiting 
seal life involved wear in the born ol the iw.al due to brief rubbin;?, during' 
!jtart-upa and wear on the axial fiealinp, la(;e!) ol the aeal rinp,!i. Both appear 
to be minor problems and would not be expected to limit fjoal life to 1(!3!3 than 
the 10-h, 300-'jtart requirement. 

In summary, deopite bein>', highly 'ionairive to rubfi between the no.al rings and 
runner as are most high-speed radial seals, the. current d(?sign appears capable 
of meeting the NASA life requirement of 10 h and 300 starts. 

3.6.4 Leji^kage Rates 

Low helium leakage rates are a very important aspect of the seal dfisign. Thi; 
leakage rates of the 5Q-mm Rayleigh-3tep design were found to be very low and 
represent a significant improvement over currently used designs. Figure 2-8 
provides an overview showing a maximum flow envelope for each seal set as a 
function of pressure drop. Each flow curve represents the higher of the f.wo 
seal (inboard or outboard) flows at a given operating point. All speeds and 
pressure conditions included in the testing are covered except for 7,ero speed. 
The highest fl.iw recorded during steady-state testing was 0.0026 kg/s 
(33 scfm). Typically, the flows ranged between 0.001 and 0.002 kg/s (13 and 
25 scfm). While the leakage flows were quite low, the govervi'iig relationships 
proved to be very complex. Thus, given the number of test conditions repres- 
ented, the reader is cautioned not to draw any cone L u; ions beyond simply 
establishing the general range of flows. 

Figure 2-9 shows a slightly more simplified overview. Again, the maximum flow 
is given for each seal set. However, in this c,?se, only test points at shaft 
speeds of 4712 rad/s (45,000 r/min) are given. While it is not the highest 
speed tested, it is the highest speed at which a wide range of supply pres- 
sures were applied to each seal set. Most of ihe flows were in the 0.0010 to 
0.0016 kg/s (13 to 20 scfm) range. Seal set Nos. 2 and 3 had maximum pressure 
drops of 1365 and 1250 kPa (198 and 181 psi) Extrapolating to a pressure 



3-32 



drop •*•■ J kPa (200 psi) results in expected flows of 0.0019 and 0.0012 kg/s 
(2A " J scfm). Because the radial flow clearances will go down as speed 
increas^.s due Co centrifugal growth of the runner the extrapolated flows could 
also be .egarded as conse-vative estimates of the flows that would have 
occurred at the full design conditions. 

3.6.5 Parametric Effects 

The testing uncovered a number of relationships among the various system 
parameters. Some of these were very logical and supportive of the design 
analysis, while others pointed out new insights that need to be included or 
more heavily weighted in the design. Because the system proved to be very 
complex, primarily because of strong parametric interaction both from within 
the system and with the surrounding environment, the measurements though fully 
adequate to verify the main performance variables (supply pressure, speed, 
leakage, etc.) were not sufficient to explain all aspects of the system behav- 
ior. With these considerations in mind, the following sections discuss the 
principal system variables in terms of: 1) how they affected or were affected 
by system performance, particularly leakage flow, 2) how their behavior corre- 
lates with the- design analyses, and 3) what emphasis, both experimental and 
theoretical, should be placed on them in future designs and studies. The 
discussion includes: 

• Seal flow path 

• Environmental interaction 

• Seal pressure 

• Shaft speed 

• Seal clearance 

• Seal temperature. 

3.6.5.1 Seal Flowpath . Most of the discussions of seal flow so far have 
assumed that all of the helium flow goes through the annular space between the 
seal ring and the runner. This is not necessarily the case. A second flow 
path exists across the radial sealing land of the ring and the mating surface 
on the seal housing. If either surface is not flat or becomes distorted, a 
flow area will exist and flow will take place. Likewise, if for any reason the 



C -5^ 

3-33 



ring Lifts off slightly, a flow will taka place. In either case, the flow 
instrumentation would not have been able to distinguish between the normal 
flow through the radial clearance and an axial bypass flow. Any bypass flow 
would have been combined witn the normal flow. 

Bypass flows, if they occur, are likely to be fairly small and therefore prob- 
ably do not have a large effect on overall performance. However, at condi- 
tions of low flow between the seal and runner, bypass flows may cause 
substantial errors in the measurements and result in poor correlation with 
predicted results. Future work should give consideration both to predicting 
and measuring bypass flow. 

3.6.5.2 Environment Interactions . The system surrounding the seal rings and 
runner had a very strong effect on the leakage rates. Moreover, effects were 
not the same for all of the seals even though the principal test conditions 
(supply pressure, speed, etc.) may have been the same. Differences occurred 
in two areas: 1) between the inboard and outboard seal rings, and 2) between 
seal set No. 4 and the first three seal sets. Subsection 3.2 describes the 
mechanical differences of both. 

The principal differences between the inboard and outboard seals were the 
temperatures of the seal rings and the corresponding sections of the runner. 
The main effect was on the runner for Che first three seal sets. The inboard 
face was directly exposed to the LN2 in the inboard drain cavity whereas the 
outboard face was expossd to the outboard drain cavity containing only helium. 
This resulted in the runner taking on the shape of a truncated cone due to a 
net contraction of the inboard end. The effect on the rings was somewhat 
less. Heat was conducted out of both rings through the axial sealing lands 
into the adjacent housing. Because of the temperature difference across the 
housing, the inboard ring was slightly colder and therefore contracted more 
than the outboard ring. This was predicted by the thermal analyses and veri- 
fied by measurements of the seal ring temperatures. Typical data is shown in 
Figure 3-10. Measurements were not possible on the runner. The net effect of 
the differences in thermal contraction between the inboard and outboard seals 
(and runner) was that the inboard clearances tended to increase more than 



3-34 



those on Che ouLooard as the unit cooled down to operating temperature. This 
permitted generally larger flows through the inboard seal ring. 

The physical changes between the fourth seal and the previous three are 
described in Subsection 3.2. The major effect was that the labyrinth seal 
prevented the draining of LN2 from the adjacent bearing from contacting the 
end of the runner and the inboard side or the seal housing. This resulted in 
two changes: 1) the thermal contraction of the runner was less on both ends, 
and 2) the coning effect was greatly diminished. Similar effects occurred 
with the seal rings. With the seal housing better isolated, its temperature 
also went up resulting in less heat transferred from the seal rings and higher 
ring temperatures. The overall result of the addition of the labyrinth seal 
was higher and more uniform seal ring and runner temperatures, e.g., less 
difference between the inboard and outboard seals and less change in radial 
clearance due to tester chill-down. The latter is illustrated by comparing 
the temperatures in Figure 3-10 and those in Figure 3-50. The latter accounts 
for the comparatively low flows measured on seal set No. 4, despite its larger 
room temperature clearances. Overall, the changes were beneficial because 
they reduced the effects of the surrounding system allowing better control 
over seal operating parameters. Future developmental tests should give care- 
ful consideration to achieving good isolation. Tests designed to evaluate the 
seal's design in a specific application, e.g. the LOX turbopump, must simulate 
the environmental interaction as closely as possible. 

3.6.5.3 Seal Pressure . Both supply and drain pressures were key independent 
variables controlled during Che testing and of obvious importance in determin- 
ing helium flow rates. While various supply pressures were applied, the 
inboard drain pressure was held fairly constant. Thus, for the inboard seal, 
the supply pressure also determined the pressure drop. The outboard drain was 
set at several different values resulting in different pressure drops and, 
therefore, flow rates at the same supply pressures. 

Experimentally evaluating the effects of supply pressure or pressure drop was 
very difficult for the first three seal tests because the operating clearances 
were not accurately known. Since substantial changes in the clearance were 
known to have taken place, it was impossible to experimentally separate the 



3-35 



effects of the pressure and clearance, at Least for the bulk of the testing. 
To provide a rational (if only first order approach), a comparative analysis 
was conducted which consisted oL computing theoretical flow-pressure curves 
at different clearances and comparing them with the measured flow-pressure 
curves. This automatically provided an estimate of the actual operating 
clearances. The data taken during seal set test No. A included good measure- 
ments of clearance on the outboard seal and thus a much better basis for anal- 
ysis. 

At low speeds most of the experimental flows from the inboard seals of seal 
set Nos. 2 and 3 were found to increase in the same manner as theoretical flow 
at constant clearance. For these data, the flow increased solely as a func- 
tion of supply pressure. At high speeds, the flow was found to be flatter with 
higher pressure points corresponding with smaller theoretical clearances than 
lower pressure points. 

Generally as the speed increased, the indicated theoretical clearance 
decreased. While some of this was clearly due to the effect of centrifugal 
growth causing the runner to increr'ie in diameter, the indicated decrease was 
more than was expected due to growth alone. This suggested that other effects 
are also taking place. The three effects are illustrated in Figure 3-82a. 

The outboard seals of set Nos. 2 and 3 showed a quite different flow behavior 
as illustrated in Figure 3-82b. Generally, the flows were considerably less 
than the inboard seals. This was most pronounced at low speed where the 
inboard flows followed fairly high theoretical clearance lines. At constant 
speeds, the outboard seal flow curve stayed flat as pressure increased indi- 
cating that the clearances were decreasing. This may have been due to the 
increase in total seal flow (due to increased inboard seal flow) which would 
have a warming effect on the runner and thus cause a net decrease in the 
already small outboard seal clearance. The second strong effect was no 
discernable change of flow or indicated clearance as speed was increased. The 
speed effect was almost the reverse of what occurred on the inboard seal. For 
the inboard seal, the decrease in indicated clearance was much greater than 
the centrifugal growth would permit; for the outboard seals there was no 
decrease. Some of the effects are baffling and strong thermal interactions 



3-36 



are suspected. The £low daCa from seaL set No. 1 was excluded from the 
discussion because only total flows were measured. 

During the testing of seal set No. 4 only one pressure was applied at each 
speed, therefore, no experimental flow-pressure curves were generated. How- 
ever, the flow decreased substantially on both seals as a function of speed, 
permitting a comparison of measured clearance with indicated clearance over a 
range of values. This showed that the actual clearances were considerably 
less than those indicated by the theoretical flow relationship (see Figure 
2-7). This further indicates that actual flows are greater than predicted 
flows at the same clearance. Two possibilities arise to explain the discrep- 
ancy: 1) the flow model needs to be modified to fully account for the condi- 
tions in the seal, or 2) the difference between actual and predicted flows 
occurred as a bypass flow. Unfortunately, the shortness of the test precluded 
exploring the behavior to any greater extent. It is important to note that 
the anomaly found with the fourth seal more than likely applies to the three 
previous tests and must be considered in evaluating their behavior 

Future work needs to concentrate both experimentally and analytically to 
better characterize the 'pressure-flow relationship and its interaction with 
seal and runner temperatures and speeds. The strong thermal effects under- 
score the recommendation of the previous section. Other items should include: 

1. Measuring seal film thickness of both seal rings using embedded 
probes. This is clearly the most satisfactory approach. 

2. Measuring the bypass flow. Even a rough measurement would be useful. 

3. Measuring the drain flow from both seals. This would provide a check 
on the overall accuracy of the flow measurements. 

3.6.5.4 Shaft Speed . Shaft speea was another carefully controlled independ- 
ent system variable. Speed has a major effect on the stiffness and damping 
properties of the Rayleigh-step part of the seal ring and hence the ability of 
the seal ring to maintain a centered position and good dynamic behavior. Its 
effect on seal leakage flows, however, is entirely indirect and theoretically 



3-37 



consists only of causing the runner to grow in diameter because of centrifugal 
forces which cause a decrease in seal clearance and Leakage flow. 

The effect of decreasing clearance with increasing shaft speed was more or 
less borne out during the testing. Data from seal set No. 1 (combined flows) 
and the inboard seals from seal sets Nos. 2 and 3 clearly showed decreases in 
indicated clearances as speed increases. The outboard seals of seal set 
Nos. 2 and 3 did not show the clearance decrease which indicates the presence 
of an opposite influence. 

The testing of the fourth seal also showed the effect and since the outboard 
seal clearance was accurately measured permitted a direct comparison with the 
predicted behavior. Figure 3-53 showed the measured decrease in clearance 
matched the predicted change. 

3.6.5.5 Seal Clearance . This proved to be one of the most elusive parameters 
with an accurate measurement not being achieved until the fourth seal test. 
Clearance is important both in determination of the stiffness and damping 
properties of the seal rig and the leakage flow rates. 

Operating clearances were found to be very sensitive to several factors 
including speed, environment effect, and flow effects. The second and third 
are entirely thermal effects and more difficult to fully characterize. The 
environmental influence from the nearby cryogenic tested section, cause the 
clearances to increase by decreasing the temperature of both the seal ring and 
runner. The clearance increase results from the larger expansion rate of the 
runner material. The helium flow, being much warmer than either seal parts, 
has the opposite effect causing the temperatures to increase and the clearance 
to decrease. 

Since leakage flow and seal clearance are so closely related, the discussions 
of the pressure-flow characteristics given in Subsection 3.6.5.3 also provide 
direct insigh.s into the clearance behavior. In summary, these arrangements 
indicated that the inboard seal clearances of seal set Nos. 2 and 3 were 
insensitive to pressure and flow and decreased as speed increased. The 
outboard seal clearances of seal set Nos. 2 and 3 decreased as pressure 



3-38 



increased and were insensitive to speed changes. Both of these effects were 
actually caused by the concurrent thermal changes. The seal data of seal set 
No. 4 were too limited to estalilish clearance sensitivity to pressure, howev- 
er, did disclose that both seals showed the predicted clearance (based on 
actual clearances) as speed increased. 

One additional consideration worth noting is the effect of the manufactured 
clearances and the initial chill-down. For seal set Nos. 2 and 3, the manu- 
factured clearances of the inboard seals were smaller than the outboard seals 
by '^30% (see Table 3-1). However, due to initial chill-down, the inboard seal 
clearances become larger than the outboard seals. This condition resulted in 
the inboard flows being generally larger than those of the outboard seals. 
The larger flow of the inboard seal was thus likely to have a stronger thermal 
effect both on its own clearance and the clearance of tne outboard seal than 
the much smaller flow from the outboard seal. Along the same lines, the smal- 
ler clearance of the outboard seal makes its flow more sensitive to a given 
change in its clearance. Both effects tend to enhance the likelihood of the 
inboard seal flow having a strong effect on both the clearance and flow of the 
outboard seal. 

The manufactured clearances of the No. 4 seals were larger than most of the 
previous seals. However, after the initial chill-down, its indicated clear- 
ances dropped below those of seal set Nos. 2 and 3. This showed the dimin- 
ished environmental effect re.'sulting from the addition of the labyrinth seal, 
and the better thermal isolation it caused. Because supply pressures were 
held constant during the run, the tendency for the inboard seal flow to 
strongly affect the outboard seal could not be established. 

The arguments again underscore the need for good clearance and flow measure- 
ments. Also, given the strong system interactions which effect clearance, 
future efforts should incorporate an extensive thermal analyses. 

3.6.5.6 Seal Temperature . The importance of both seal and runner temper- 
atures in determining clearance and leakage flows has already been estab- 
lished. Runner temperatures are probably more important because of the higher 



3-39 



expansion rate of Inconel. However, because the runner rotates temperature 
measurements are virtually impossible and were not attempted during the 
program. At best, seal ring temperatures provide only part of the desired 
relationship and help to establish trends. 

One effect common to all of the seals was the extremely low seal temperature 
that resulted at very low helium" flow rates. Temperatures of -70 to -gO^C 
(-94 to -130°F) were recorded. This is not unreasonable because the only 
source of heat is that transmitted through the seal housing from the outside 
air. For the first three seal sets, the rate of change of seal temperature 
with supply pressure was very high at low pressures and gradually tapered off 
at intermediate and high pressures. 

The No. 4 seal data did not permit a parallel assessment. The temperature 
data given in Figure 3-50 did show, however, that with no flow (which was the 
case after the tester shut down), the seal temperatures did reach -70°C 
(-94''F). However, when a small flow was present, the temperatures quickly 
rose to -10 to +10''C (14 to SO^F). ^See the -100 to -60 min period on Figure 
3-50.) Figure 3-82c illustrates the effects. The effects of speed on seal 
temperatures were significant in most cases. Seal set Nos. 1 and 3 showed a 
roughly parallel downward shift of the temperature supply pressure curves as 
speed increased (see Figure 3-82d). The temperature data of seal set No. 2 
did not show a clean trend. Seal set No. 4 also exhibited a temperature 
decrease as speed increased. 

A last observation worthy of mention is the effect illustrated in 
Figures 3-82e and 3-82f. This was described in the latter part of 
Subsection 3.2.2. Figure 3-82d shows several curves of flow and temperature 
versus pressure drop across the outboard seal of seal set No. 2. The curves 
represent high, medium, and low outboard drain pressures and show that while 
as outboard seal flow increased steadily as the pressure drop increased, the 
corresponding seal temperature curves showed major discontinuities. Figure 
3-82f plots the same flow and temperature data plus the total flow from both 
seals versus supply pressure. The figure also shows the temperature curves 
falling much closer together. Since the inboard seal drain pressure was the 
same for all the runs, its flow, and hence the total flow from both seals, 



3-40 



function of supply pressure. Therefore, total flow appeared to be a much 
greater factor in determining the outboard seal temperature than was the 
outboard seal flow. 

3.6.6 Dynamic Performance 

The dynamic performance of the seals tested proved to be very good. However, 
the testing was not designed to evaluate this aspect in any particular fash- 
ion, therefore, no controlled excitations were applied. Neither were the 
operating conditions intentionally changed to require the seals to run in a 
region where self-excited motions were predicted to take place; i.e., the low 
friction region. In fact, quite the contrary, drain pressure requirements 
resulted in the seals operating almost totally in the high friction region of 
the operating map where dynamic motions are heavily retarded by substantial 
friction forces. 

Seal runner whirl orbit diameters were generally in the Q.005 Co 0.008 nnm 
(0.0002 to 0.0003 in.) range except for some of the high speed runs where 
orbit diameters reached 0.010 'to 0.013 mm (0.0004 to 0.0005 in). For most of 
these runs, the seals remained motionless which is the predicted response 
considering the operating region. It is also the preferred response since the 
whirl orbits of the sizes described are small in relation to the operating 
clearances generally observed. 

Seal motions did arise on several occasions. The principal occurrence was 
with the inboard seal of seal set No. 3. The seal ring developed an in-phase, 
generally elliptical orbit. The orbit was observed at all three test speeds, 
3665, 4188, and.4712 rad/s (35,000, 40,000, and 45,000 r/min) and became larg- 
er in amplitude as the supply pressure was increased. The maximum orbit diam- 
eter was approximately 0.025 to 0.030 mm (O.OOl to 0.0012 in.). Since the 
supply pressures and shaft speed placed seal operation clearly in the high 
friction region, the motions are baffling. Details of the occurrence are 
provided in Subsection 3.3.2. 

In summary, the tests did show that the seals were generally well behaved 
dynamically, but in view of the lack of specific dynamic testing and the 



3-41 



occurrence of large whirl, it is suggested that future studies give particular 
attention to this area^ Analytical studies should consider how a seal ring 
might respond in the presence of a significant bypass flow. 

3.6.7 Material Co nsideration s 

All three seal failures were due to rubs between one of the seal rings and the 
runner. While each of these was precipitated by a different mechanism, the 
end result was unacceptable damage to both parts. Two combinations of materi- 
als were tried. The first consisted of P5-N formulation of carbon graphite by 
Purebond for the seal ring against a tungsten carbide coating using a silicon 
carbide binder on an Inconel 718 runner. This was used for the first three 
seal tests. Two failures occurred with this combination resulting in substan- 
tial wear of both the seal ring and runner. Also, numerous surface cracks and 
some delamination of the carbide coating occurred on the runner. Very high 
temperatures had been generated at the rubbing interface and were responsible 
for the cracking and rapid deterioration. The second combination which was 
used for the fourth seal set did not work any better. It consisted of the same 
seal ring material, however the runner had an electrolized surface with 
Inconel 718 again as the base material. The failed parts again showed 
evidence of rapid wear and high temperatures. While no cracks were observed 
on the runner surface, the grooving and wear were no more acceptable than the 
damage of the previous runners. 

Given these results, it is clear that additional work needs to be done to 
identify or develop material combinations that are more suitable. While good 
strength properties are necessary and important, good rubbing properties at 
both low and high temperatures are key. Low coefficients of friction and high 
thermal diffusivity are very important. Other necessary properties need to be 
identified. 

Another aspect that needs to be considered is the match-up of coefficients of 
expansion of the seal ring and runner materials. The base materials used in 
the testing have substantially different rates, 10.8 )im/m-°C 
(6.0 liin./in.-°F) for the Inconel 718 runner and 4.3 Jim/m-^C 
(2.4 Jiin./in.-°F) for the P5-N carbon graphite seal rings. The much higher 



3-42 



rate for the runner resulttj in a tendency for the runner fn rapidly >?row into 
the seal due to the heat generated during the initial sta>'ei of a cub. This 
would aggravate a partial rub and, because the operating clearances are very 
small, would quickly result in a progression to a full rub. Material combina- 
tions with closer expansion rates or in which the seal ring material has high- 
er rates than the runner material would help to alleviate the problem. 



3-43 



TABLE 3-1 



SEAL DIMENSION SUMMARY 



Seal Sat 
Number 



Design 
Dimensions: 



I 



Serial 
Number 



Outboard Ring 
Inboard Ring 
Runner 



Mating Surface 

Di^ameter 

(mm) 



Radial 

Clearance 

(mm) 



50.028-50.033 
50.020-50.025 

50.G02-50.010 



0.009-0. 015 
0.015-0.01 1 



Outboard Ring 
Inboard Ring 
Runner 

Otb. Sur. 

Inb. Sur. 



108302 
108306 
078303 



50.030-50.033 
50.020-50.024 

50.UtiQ-50.004 

50. tjaO-50.004 



0. 013-0. 0T7 

0.008-0.012 



Mating Surface 

Diameter 

(mm) 



50.079-50.086 



49.992-50.028 
49.995-50.002 



Radial 
Clearance 



0. 025-0. C47 



Outboard Ring 
Inboard Ring 
Runner 

Otb. Sur. 

Inb. Sur. 



108304 
108308 
078302 



Outboard Ring 
Inboard Ring 
Runner 

Otb. Sur. 

Inb. Sur. 



50.029-50.033 
50.Q23-50.025 

50.ii09-50.01 1 
50.009-50.01 1 



0.009-0.012 
0.006-0.008 



50. 038-50. 04J 
50.033-50.035 

50.029-50.011 
50.009-50.01 1 



108305 
108301 
078302 



Outboard Ring 
Inboarj Ring 
Runner 

Otb. Sur. 

Iiib. Sur. 



50. 029-50. 030 
50.020-50.025 

50.G4i9-50.01 1 
50.ls09-50.01 1 



0.009-0. on 

0.004-0.008 



103301 
108302 
078304* 



50.'i29-5G.033 

50.020-50.025 

49.937- 
49.992-49.995 



0.016-0.018 
0.013-0.016 



50.038-50.089 

50.030-50.033 

50.000-50.025 
50.005-53.008 



0.G01-Q.Q44 



*MTI 



» ^ 



1^1 



TABLE 3-2 



ACCELERATION RUN PERFORMANCE DATA - SEAL SET NO. 2 



I 



Run No. 



1-5 

6 

7-1 1 

12 

13-17 

18 

19-23 

24 

25-29 

30 

30a* 

31-35 

36 

37-46 

47 

48-57 

58 



Acce I erat ion 

Rate 

H - High/L - Low 



Ma X i mum 
Snaft Speed 

(rad/s) 



He I ium SuppI y 

Pressure iPlIJ 

(kPa. Abs. J 



H 

L 
H 

L 
H 
L 
H 
L 
H 
L 
L 
M 
L 
M 
L 
H 
L 



4. 188 



931 



Outboard Drain 

Pressure CP2) 

(kPa. Abs. 5 



5i: 



Outboard Seal 

Fio« 
tkg/s X 



JO^) 



Inboard Seal ' 
FIoh 

Ckg/s X 30*) 



s 


* 


8 


5 


8 


9 


9 


2 


9 


3 


6 


7 


6 


5 


7 


6 


8 


8 



9.1 

7.8 

7.4 

6.3 

6.2 
17.2 

15.4 

II. ■> 

8.8 



*Run 30a was a repeat of k^n jsj, after a brief shutdown 



:fian9e chart paper. 



TABLE 3-3 



ACCELERATION RUN PERFORMANCE DATA - SEAL SET NO. 3 



I 





Accel erat 1 on 


Max imum 


Hel lum Supply 


Outboard Drain 


Oi-tboard Sea) 


Inboard Seal 




Rate 


Shaft Speed 


Pressure (P12) 


Pressure (P2) 


Flow 
Ckg/s X 10^) 


Flow 
{kg/s X 10^) 


Run No. 


H - High/L - Low 


(raa/s) 


(kPa, Abs.) 


(kPa. Abs.) 


1-5 


H 


4.712 


1 .069 


517 


_ 


_ 


6 


L 












10.4 


'1.5 


7-1 1 


H 






1 r 






- 


- 


12 


L 












9.6 


13.1 


- 








1 .206 






10.8 


14.1 


13-17 


H 












- 


- 


18 


L 












9.6 


17.3 


19-23 


H 












- 


- 


24 


L 






^ 






9.2 


25.6 


- 








1 .344 






9.2 


27.9 


25-29 


H 












- 


- 


30 


L 












8.8 


28.3 


31-35 


H 












- 


- 


36 


L 






T 






8.8 


27.5 


- 








1 .482 






9.6 


31.2 


37-41 


H 














- 


- 


42 


L 














8.8 


31.1 


43-47 


H 














- 


- 


48 


L 














8.9 


31.3 


48a* 


L 


' 


' 


T 


U 







*Run 48a was a repeat of Run 48. after a brief shutdown to change reels of magnetic tape. 



TABLE 3-4 



SEAL INSPECTION SUMMARY 



U) 

I 



Seal Set 
Number 


Outboard Ring 
S/N 


Inboard Ring 
S/N 


Runner 
S/N Material 


Pre-Test 
Condi t ion 


Post-Test Condition 


1 


106302 


108306 


078303 
Inconel 718 with 
tungsten carbide 
coating. 


Al 1 new parts. 


Both seal rings intact. Outboard 
ring had moderate surface wear. 
Inboard had no wear. Runner had 
numerous heat cracks and some 
delamination of coating. 


2 


108304 


108308 


G78302 
Inconel 718 with 
tungsten carbide 
coat ing . 


Al 1 new parts . 


No we&r on seal rings nor runner. 
Very slight deposit of carbon 
graphite on axial seal lands of 
seal housing. 


3 


108305 


108301 


078302 
Inconel 7l6 with 
tungsten carbide 
coat ing. 


New seal rings; 
runner from pre- 
vious test; no 
evidence of wear. 


Seal rings intact. Outboard ring 
had moderate wear while inboard 
ring showed none. Heat cracks 
and some delamination on runner. 


4 


108301 

Integral 

Probes 


108302 


078304 (MTI) 
Inconel 718 with 
el ectrol i zed 
surface . 


Al 1 new parts. 


Outboard ring intact. Inboard 
ring had radial crack and showed 
moderate wear. Runner surface 
gal 1 ed and Trn. 



a 
a 



a. 



5 
1/1 

Ul 

a 
a 



in 



X 



1200 



UOO 



:ooo 




J 



r 



14-flUN NO. 3' 



•4- 



40 50 

TIME imfn) 



SO 



70 



ao 



b 



90 



Fig. 3-1 Pressure History - Seal Set No, 1 



SOOOi 



5300 



56Q0t- 



H 



SEAL SET NO.l 



520O1 



T 



c ,^ 1 .•< — — 



73 



a 



f 



-.ii. 






J2C31 

J 
I 

350 






JMUL 



3z; 



^ 



•PUN jNO.X- 



=ftL 



i 



-flUN Nb.2-|N 



-•"■■i^i 



J 




! t 



i i 



I I 

-ri- 
■■ I 



X-flUN N0.3-ft 



10 



20 



30 



40 SO 

TIME (Bifn) 



60 



70 



80 



90 



), 1 



Fig. 3-2 Speed History - Seal Set No. 1 



3-48 



r4 



T?r 



I 

VO 



a 



a. 



in 
in 

UJ 

(X 
a 



13 

in 

3 




1000 



2000 



3000 4000 5000 

SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Fig. 3-3 Operating Map for Seal Set No. 1 



O 



120 



'T 



100 - 38- 



80- 



60- 



0) 

k. 

3 
CS 

Q. 

E 

0) 



^ 40 



CO 



20 



J 



°C 
49-1 



rad/s rpm 



27- 



16 



4- 



-7 -J 



-18-> 




Speed 



Onset of Fal 



ure 



Outboard Sea 
Temperature 



Inboard 






Seal 



Temperature 



Time 



-4eU- 



6282 



-5235 



-4188 



-3141 



-2094 



-1047 







r 60.000 



-50.000 



-40.000 ■§ 

(D 
Q. 
CO 

© 

-30.000 I 



-20.000 



10.000 



L-0 



« . * ( i t; 



Fig. 5-4 Failure of No. 1 Outboard Seal Ring 



851151 



sr\ 



.007t 



.006 



.005 



517.0 

37.1 

LIS* 

50.0 

0.0 

1.0 

2.07 E- 

l.ti 

4.003 

ZERO 



VI .004-- 






.003' 



.002 



.001 




OOUNSTRCAM PMCSSUPE (KPa) 

UfiSTRCAH TEMI>ERATURe((lt{| CI 
SEAL LENfiTH(M> 
SEM. DIAMETER (m) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (P«-*) 
AIIAIATICCOCFFICIENT (ga«M> 
MOLECaAR WEIGHT 
SPEED (rad/a) 
THEOAETICAL DATA 
EXPERIMENTAL DATA 
(HAUF TOTAL FLOW) 
SEAL SET NO.l 



100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 

PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-5 Zero Speed Flow versus Pressure Drop - Seal Set No. 1 




100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-6 Zero Speed Seal Temperatures versus Pressure Drop - Seal Set No. 1 



3-51 



.005r 



.004 - 



.003 



I 

LP 



(A 



3 
O 



.002 - 



,001 - 



517.0 

37.8 

1.956 

50.0 

0.0 

1.0 

2.07 E- 

1.66 

4.003 



DOWNSTREAM k^ESSURE (KPa'. 
UPSTREAM TEMPERATURE (deg C) 
SEAL LENGTH (m) 
SEAL DIAMETER (mm) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
08 VISCOSITY (Pa-s) 

AOIABATIC COEFFICIENT (gsHUM) 
MOLECULAR WEIGHT 

— THEORETICAL DATA 

— EXPERIMENTAL DATA 

(HALF TOTAL FLOW) 
SEAL SET NO.l 




100 



200 300 400 500 
PRESSURE DROP ACROSS SEAL (KPa) 



600 



700 



Fig. 3-7 Flow versus Pressure Drop - Seal Set No. 1 




100 



300 300 400 SOO 
PRESSURE DROP ACROSS SEAL (KPa) 



700 



Fig. 3-8 Inboard Seal Temperature versus Pressure Drop - Seal SeE No. 1 



S754.5 (rM/») 



100 



200 300 400 500 
PRESSURE DROP ACROSS SEAL (KPa) 




jsr-i 



600 



700 



Fig. 3-9 Outboard Seal Temperature versus Pressure Drop - Seal Set No. 1 

3-53 



I Or 



I 






LU 
CX 

H- 

o. 

LU 




200 300 400 500 
PRESSURE DROP ACROSS SEAL (KPa) 



600 



700 



Fig. 3-10 Inboard and Outboard Seal Temperature versus Pressure Drop - Seal Set No. 1 



£ 



1500 

1400 

1300 

1200 

UOO 

1000 

900' 

800 

700' 

600 

SOO 

400 



SEAL Ser NO. 2 






t— l-=....-4».--"i- 




20 30 



40 50 60 70 

TIME (mfn) 



80 



90 100 110 120 



Fig. 3-lJ First Day Pressure History - Seal Set No. 2 



6000 
6800 
5600 
5400 
5200 
5000 



, " I "T ' ! r i - 1 


1 ! ; 


~", — "•■*- --r- — 


' ■ 


I 




j j 


SEAL SET NO. 2 


r r""""' ■ 


t t 


i 1 


i, 1 . ' i i j 


r ! ' , i i i ' 




50 60 70 
TIME (Mfn) 



120 



Fig. 3-12 First Day Speed History - Seal Set No. 2 



3-55 






<« 
a 



t/l 



> 
-I 

a 

% 
I/) 

5 




10 30 30 40 SO 60 70 80 90 100 110 120 130 

TIME (mln) 



Fig. 3-13 Second Day Pressure History - Seal Set No. I 



6000 

5600 

5600 

5400 

S?00 

5000 

^ 4800 

g 4600 

I. 

■" 4400 

Q 

^ 4200 
4000 
3800 
3600 
3400 
3200+- 



...■■i u... t.-. 

■ i ; i- 



• i ;- !....-+•■;•• 

■•♦ ■. - U" 



3000- - 



..i 



—r-- 



"!-■- 



SEAL SET NO. 2 



.\.. -..4-. 

4- • ' i i..- 



+= 



"-!■"-— I" 



^' 



XZ± 



Liznznz! 






— -..+... 



i™.-..».i . 

•■— -i 4- 



1 
— ,. 



.. 1 



■cm 






—.B^**..!tS.:.l. 



..U....4. 



•t 









1 I j ■ 

,-4 — U 1. 



••<-*■ 






..U. 



,.._4-.j.._,.,J 



-U 



..-.+...1.. 



I i 



!SLJi^72 



4-- 



10 20 30 40 SO 60 70 80 90 100 110 120 130 140 ISO 

TIME (Mfn) 



Fig. 3-14 Second Day Speed History - Seal Set No. 2 



3-56 



I " 




8 9 10 U 12 13 14 IS 16 17 18 
TIME (Wtn) 



Fig. 3-15 Third Day Pressure History - Seal Get No. 2 



6000 
5800 
5600 
5400 
5200 
5000 

^ 4aoo 

I/) 

o 4600 

u 

■^ 4400 
a 

S 4200 
4000' 
3800 
3600 
3400 
32C0 
300O 



- ** ■'"ii " •*• "'^ '*■■ 



SEAL SET NO. 2 

■■■■^ ^■.:.^..4...J:;:..i '.f. 



n.i„.„.4...~| — |..™X--.4 — I — f 



„_4-. 















t«/H»f««t ■•i««*^a«nicn^B>M 



.l.„.. 



-♦_.-..4— *- 



" ' ' " * r ~* ^....~.+."-»~^.-.. ^.™j«_„^™.-...j j.„ »._«_.,.. ..)._.-.*_. ...4, , 



iznx 



..»4.. ».L«. 






"! — r 



t 



_9!l!!l.®i.5. 



"i — r — ■< 



tl 



~t- 



'. — -t 
•1 — ' 



z 



~.u 

,-_i 



8 9 10 11 12 13 14 IS 16 17 18 19 20 
TIME (Mfn) 



Fig. 3-16 Third Day Speed History - Seal Set No. 2 



3-57 



I 

00 






a. 



UJ 
DC 

r> 
a 



a 
in 



UJ 




1000 



2000 



3000 4000 5000 

SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Fig. 3-17 Operating Map for Seal Set No. 2 



.00?T 



.00«- 



.005 



ui .004- 



01 
X 



31 
O 



.003- 



.002 



.001 





r-r- •.,,-,-.-^,-,...,-.:-.T - = :=..- '- ' 


9W.a 


OOWMtMAM MCSSUMt (K^r« 


17.1 


\JHmm TEMMERATUM (dtg C) 


I.MS 


SCM. LEMTH (M) 


10. ft 


tiM. OIAMCTIN (M) 


o.t 


rCCCNTRICITV RATIO 


1.0 


OISCHARCE COfFriCIENT 


l.iJ 6-0« 


VISCOSITY (R»-*) 


t.i« 


AOIAIATIC COEFFICIENT (gaaM) 


4.003 


MOLECULAR WEIGHT 


ZERO 


SREEO <ria/*) 
THEORETICAL DATA 




■•■*•"«••■'> 


EXMRIHCNTAL OATA 




SEAL SET N0.2 




INCOARO SEAL 



^^siE:£-:-^r^-*a 




100 200 300 400 500 SOO 700 800 900 1000 UOO 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-18 Inboard Seal Flow versus Prasstiro Drop. Zero Speed - Seal Set No. 2 



.O07r 



.006- 



.005- 



ifl .004- 



oi 



•a 
o 



.003 



.002- 



.001 - 



..>,«.. 



517.0 

37.1 

I.MS 

SCO 

0.0 

1.0 



OOUNSTREAM PRESSURE (KPl) 
UPSTREAM TEHPCRATURC ((MO C) 
SEAL LENGTH (M) 
SEAL DIAMETER (m) 
ECCENTRICITY RATIO 
OISCHARCE COEFFICIENT 



2.07 E-Ot VISCOSITY tP«-«) 



1.6t 

4,003 

Ztfu 




AOIAIATIC COEFFICIENT (g«MM) 

NaECULAR UEICHT 

SPEED (rid/a) 

THEORETICAL OATA 

EXPERIMENTAL OATA 

SEAL SET NO. 2 

OUTIOARO SEAL 



100 200 300 400 SOO 600 700 800 900 1000 UOO 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-19 Outboard Seal Flow versus Pressure Drop, Zero Speed - Seal Set No. 



3-59 



o 





Dl 




01 




r> 




^-^ 




Ui 




a: 




3 




1- 




<£ 




QC 


LO 


Hi 


o 


a 



-10 - 



-20 - 




-30 - 



-50- 



400 500 600 700 800 900 1000 1100 1200 1300 
PRESSURE DROP ACROSS SEAL (KPa) 



1400 



Fig, 3-20 Zero Speed Seal Temperatures versus Pressure Drop - Seal Set No. 2 



.007r 




300 400 500 600 700 
PRESSURE DROP ACROSS SEAL (KPa) 



1000 



Fig. 3-21 Inboard Seal Flow versus Pressure Drop - Seal Set No. 2 






a 

a 



-10 



-20 



-30 - 



.40- 



-50- - 



-60- - 



•70 



-80' 
-90 
■100 




4712 |(rad/a> 




I 



- SEAL SET IW.2 
INBOARD SEAL 
OOUNSTREAM PRESSURE - 517.0 (KPa) 



100 200 300 400 500 600 700 
PRESSURE DROP ACROSS SEAL (KPa) 



800 



900 



1000 



Fig. 3-22 Inboard Seal Temperature versus Pressure Drop - Seal Set No, 2 



3-61 



,007 



.006 



.005 - 



517.0 

37.8 

1.956 

50.0 

0.0 

1.0 

2.07 

1.66 

4.003 



E-08 



u> .004f- 



DOWNSTREAM PRESSURE (KPa) 
UPSTREAM TEMPERATURE (deg C) 
SEAL LENGTH (nm) 
SEAL DIAMETER (mm) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (Pa-S) 
ADIABATIC COEFFICIENT (gama) 
MOLECULAR WEIGHT 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 2 
OUTBOARD SEAL 



I 




i 



*.. 



100 



200 



300 400 500 600 700 
PRESSURE DROP ACROSS SEAL (KPa) 



800 



900 



1000 



Fig. 3-23 Outboard Seal Flow, 517 kPa Drain Pressure - Seal Set No. 2 



IjO 
I 

ON 



,007t 



.006 



.005 



M .004 



CK 



3 
O 



.003 - 



.002 



.001 



310.2 
37. • 
1.95C 
56. 

a.o 

1.0 

2.07 E-OI 

1.66 

4.003 




OOWNSTREAH PRESSURE (KPa> 
UPSTREAH TEMPERATURE (d«0 C 
SEAL LENGTH (m) 
SEAL DIAMETER (m) 
ECCENTRTCITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (Pa-s) 
ADIABATIC COEFFICIENT (ga 
MOLECULAR WEIGHT 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 2 
OUTBOARD SEAL 



400 500 . 600 700 800 900 lOQO 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-24 Outboard Seal Flow, 310 kPa Drain Pressure - Seal Set No. 2 



I 



o 

0) 
ID 



a. 

3 

<t 
q: 

Q. 




100 



100 



200 



300 400 500 600 700 
PRESSURE DROP ACROSS SEAL (KPa) 



800 



900 



1000 



Fig. 3-25 Outboard Seal Temperature, 517 kPa Drain Pressure - Seal Set No. 2 



-30 - 



u 





cn 




01 




TD 




' — ■ 




UJ 




a: 




3 




. h- 




•I 




o: 


LO 


^ 


a^ 
m 


Q. 




-60- 



-70 - 



-90- 



-100 



400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-26 Outboard Seal Temperature, 310 kPa Drain Pressure - Seal Set No. 2 



I 

0^ 



Or 



10 - 



20 - 



-30- 



o 



^ -40 



LU 

a -50f 

t- 
a: 



60 - 



-70 



-80 



-90- 



-1004 



SPEED 



+ 






3665 1 (:rad/a) 






4- 



4- 




SEAL SET NO. 2 
OUTBOARD SEAL 
DOWNSTREAM PRESSURE - 103.4(KPa) 



100 200 300 



4- 



4- 



4- 



4- 



4- 



4- 



4- 



■4 



400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-27 Outboard Seal Temperature, 103 kPa Drain Pressure - Seal Set No. 2 



c\ 




Fig. 3-28 



300 400 500 600 700 

PRESSURE DROP ACROSS SEAL - KPa 

Outboard Seal Temperature and Flow versus 
Pressure Drop - Seal Set No. 2 



900 



1000 



70 



3t 



3 
O 



C3t 
O 



CC 

3 

< 

a 
a. 

X 




400 



500 



600 700 800 

SEAL SUPPLY PRESSURE - KPa.abs 



900 



Fig. 3-29 Outboard Seal Temperature and Flow versus 
Supply Pressure - Seal Set No. 2 



1000 



3-67 



I 

00 



E 



in 

it 
o 



.IS 
• Mi 
.13 

.12 - 
.11 
.1 - 

.09 
.08+ 



-.06999 - 



Uj 

in 



T" 



SEAL SET NO. 2 



— J — 

t 

I 

I 

I 
— I — 



I- 






■}•■•- 



]-.... 



i 

I 
1 



I 
f- 

i 
t 
t 

• i" 

i 



I 
.i 

i 

I 

I 

I 
..j..... 

1 
i 



-i 



»•• 



... . — f... 

! 
I 

! 



+- 

i 



I- 



1 — 

I 

! 



4... 

1 



I ~ ' ! 

: ! 




0TB SEjUL VERT^I i o - ■ 



:o CD qo 



-j 

i 




\l 
i • 
.41 

• > 

• I 

• I 

I 

■ ■ 

"r< 



1»4- 



-iRUN N0i2 



; - * !j- 

ilNB SEAIJ VERT- 



'^^4 






i'lNB SEA|. HOR 
0TB SEa|. HOR- 



4^« 



-w-' 



A 



GD<S O 






'O o* 






INB SdAL VERT I j 



1 *** *4 






-•J — 

! 



4- 



4- 



50 60 

TIME (m«n) 



70 



80 



90 



100 



110 



Fig. 3-30 First Day Seal Film Thickness History - Seal Set No. 2 



UJ 

I 

U3 



in 

UJ 

z 



in 




.06999 



60 70 80 
TIME (mfn) 



120 130 



150 



Fig. 3-31 Second Day Film Thickness History - Seal Set No. 2 



,i iri i>» «jitl i . »iWliWii«M«*i» 



o 



E 



.IS 

.M 
.13 

.11 
.1 



in .09 



UJ 
CI 



.08 



*-. 06999 
u. 



.J 



m 



,05t- 

,04 
,03 
,02-h 
,01 



»^-4^iS^-2ffl 



SEAL SET NO. 2 



« r 



f • - • •» !•- - 



.j-SfTT-.}. 

» f 

j : 

i i— 

j j- 



^^- i 



I 
: 
-i — 

} 
s 






4- 



I 



;actao ': 






.^ — ._.t.. 






f 

I 

— f- 

I 



I ! ! 

-t — ^ — f 



I 



J 



5 



OUllOARD SE)(L vetlTICJfL 
, ^ j. J ^„ 




I 
! 
: 
i- 



1 

— t... 
t 



i 



ssao-i 



- — i- 






.i«ftt.l. 



r-j- 



...I- 4 — 

i : 



■I » 



Irunino.s i I 
; I ! f i 



3 : 

...4. {.^ — -.i. — 



r 1 



_., ^ — 

I : 



1 iNfldARb $EAL HoRiioNT aC bUTBOAftb sTjiLHteizojif wT 






-» T 1- 



??4Sil. 



Onboard seal vfERiitAL 



— -«• I 1 — *- — I — -* — 



i 







4— j — i — I- 



4- 



A — ^ 



4- 



I — 



..+ — t. 



... — ^._. 



i 




1 



4.. 



-! 1 1 1- 



..« — -I, 
I i 



4- 



8 9 10 11 12 13 14 15 
TIME (mfn) 



16 17 18 19 



-1 
20 



Fig. 3-32 Third Day Film Thickness History - Seal Set No. 2 






lfcss^»» 



a 



U1 

I/I 



% 

•J 




10 20 30 40 50 60 70 80 90 100 120 140 160 180 200 230 240 

TIME (mtn) 



Fig. 3-33 Pressure History - Seal Set No. 3 



6000' ' 

5800f 

5600 

5400-j- 

5200- 

5000- 

^ 4800f 

ui 

o 4600 

L. 

^ 4400+- 

o 

^ 4200- - 
4000- - 



SEAL SET NO. 3 



■•+- t" 



„4— 4- 

H — f- 



•••■i"—'f— 
••4" — 4-' 



...;~-..4- 
•••(••—■■'-■ 



••t- 7" 

-T-H" 



-T~T" 



;„..,.i- 



... i««.4.<»-.^. 



+•• 




;trx4 



3800- r- •••(• ■-4. J.-i--i 1"- 

-pA-^^MujUr'^ ! — 

3600- -'J^I^Pl-l 



:zz^^.4XIlJz: 



3400 



3200- - 



3000 



-t-t 



X 



+ 



:-f!|*l.^,.««l 



H i- 



J 



X. 



J u. 



-t— h 



■^ — I — I" 



£U9-Mii. 



4-^ 



-I- 



m 



"1 ! — 



ByHJ!<L.jidi^ 



■1 



10 20 30 40 so 60 70 80 90 XOO 120 140 

TIME (lilfn) 



160 180 



200 



220 240 



Fig. 3-3A Speed History - Seal Set No. 3 



3-71 



CO 
I 





1500 




1400 




1300 




1200 


.^^ 




(A 




ID 


1100 


a. 


1000 


i^ 




^_^ 




UJ 


900 


cc 




-) 




tri 


800 


U) 




UJ 




q: 

CL 


700 


>- 




_i 


600 


o. 




3 




in 


500 


X 




3 




n 


400 


_t 




UJ 




X 






1000 



2000 



3000 . 4000 5000 
SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Fig. 3-35 Operating Map - Seal Set No. 3 



°F 



°C 





lUU - 


JO- 




80- 


27- 


B 


Q. 

E 
o 

H 


60- 


13- 


15 
o 

£0 


40- 


4- 




20- 


-7- 




0- 


-18- 




rad/s rpm 
TT,sfi r5235 r 50.000 



4188 -40.000 



-3141 



2094 



-30.000 



Time 



-H7.5h- 



^0 



T3 
(D 
(D 
Q. 

CO 



-20.000 i 



CO 

ID 



1047 -10.000 



•-0 



Fig. i-ib Failure of No. 3 Inboard Seal Ri 



ng 



851150 



.007t 



.006- 



.005' 



ji .004 
□1 



3 

2 .003 



.00? 



.001 



517.0 

37.* 

l.lSi 

so.o 

0.0 

t.o 

2.07 E-OI 
l.i* 

«.ao3 



OOUNSTRCAH PMSSUDC (KPt) 
UKTACAN ni#CRArURC ((MO C) 
SCM. LENOTH (M) 
SCM. OI/MCTER (m) 
£CCENTRICITV RATIO 
0ISCHARC£ COCFriCIENT 
VISCOSITY (P«m) 
WIAIATIC COEFFICIENT (gaiMt) 
MOtECULAfl WEIGHT 
TI«OR£TICAt DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 3 
INIOARO SEAL 



] I 




100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL CKPa) 



Fig. 3-37 Inboard Seal Flow versus Pressure Drop - ?eal Set No. 3 



-60 



SEAL SET NO. 3 
INBOARD SEAL 
DOWNSTREAM PRESSURE 




517.0 (KPa) 



■+. 



■4- 



100 200 300 400 500 600 700 800 900 1000 1100 1200 13Q0 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-38 Inboard Seal Temperature versus Pressure Drop - 
Seal Set No. 3 



3-74 



I 



.007r 



.006 - 



.005 - 



<A .004 






.003 - 



.002 - 



.001 




100 200 300 



400 500 600 700 800 900 1000 1100 1200 13Q0 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-39 Outboard Seal Flow, 517 kPa Drain Pressure - Seal Set No. 3 



I 

0^ 



.007t 



.006 - 



.005 



M .004 - 



cm 



3 
O 



.003 



.002 - 



.001 



310.2 

37.1 

1.95« 

50.0 

0.0 

1.0 

2.07 E-Ot 

1.06 

4.003 



DOWNSTREAM PRESSURE (KPa) 
UPSTREAH TEMPERATURE (Oag C) 
SEAL LENGTH (m) 
SEAL DIAMETER (m) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (Pa-6) 
ADIAIATIC COEFFICIENT (g«HM) 
MOLECULAR WEIGHT 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET N0.3 
OUTBOARD SEAL 




400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-40 Outboard Seal Flow, 310 kPa Drain Pressure - Seal Set No. 3 



.>.-i t 



I 

-J 



.007t 



.006 - 



.005 - 



M .004 



□I 



3 
O 



.003 



.002 - 



.001 - 



103.2 

37.t 

l.«5f 

56.0 

0.0 

1.0 

2.07 E-Ot 

l.«6 

4.003 



DOWNSTREAM PRESSURE (KPa) 
UPSTREAM TEMPERATURE (dcfl C) 
SEAL LENGTH (mi) 
SEAL DIAMETER (m) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (Pa-s) 
ADIAIATF^ COEFFICIENT (gaaM) 
MOLECULAii WEIGHT 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 3 
OUTBOARD SEAL 




I 

L....... I 



100 200 300 400 500 600 700 BOO 900 1000 1100 1200 1300 1400 

PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-41 Outboard Seal Flow, 103 kPa Drain Pressure - Seal Set No. 3 



ae908 





Ln 




OJ 




o 




^-' 




UJ 




or 




J 


UJ 


1 - 


1 


•■X 


^J 


IX 


CO 


lU 



a 

5" 
h- 




100 200 300 



400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig, 3-42 Outboard Seal Temperature, 517 kPa Drain Pressure - Seal Set No. 3 



^ 



I C \ i 



IfiSit. i* 



I ^ 1 )r-iii 1 t 1 « -1 



1 I 1 ■-- -1 « J » - ' 



w 
I 



Oi 

n 



IX 
D 

I— 

•a 

a 
111 
a 



?0 



10^ 



10 



20 



30 



40 



-50 



60- 



SPEro - 3665 (rajj/s) 




+ 



100 200 300 



4- 



■4- 



SEAL SET NO. 3 
OUTBOARD SEAL 
DOWNSTREAM PRESSURE 



- 517.0 (KPa) 



4- 



4- 



4- 



4- 



400 500 600 700 800 900 1000 
PRESSUf^E DROP ACROSS SEAL (KPa) 



1100 1200 1300 1400 



Fig. 3-43 Outboard Seal Temperature, 310 kPa Drain Pre oure - Seal Set No. 3 



I 

00 

o 



0) 
o 



Ul 

»- 
•J 

or 

tj.j 
a 

Ul 

t-- 




100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 

PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 3-44 Outboard Seal Temperature, 103 kPa Drain Pressure - Seal Set No. 3 



I I ?-- -J" » ! ; 3 ! ' ' ' • 



I '- \ 



% _ ! »-^i.. \ 



» -- » I J r- -i^i '^* - ' 



u> 
I 

00 



1500 

1400 

1300 

_ 1200 

« 1100 

* 

^ 1000 

q: 

in 800 

Ul 

S 700 

>- 

600 

500 

400 

300 

200 

100 





a. 
a 

VI 

3 



UJ 



SEAL SET NO. 4 



f • 



HIGH FRICTION: REGION 












i 



SEAt TESTJ>OINJS 
o am i 



NAX0HUH DRAIN (PRESSURE 




i 

-f— 

i 
i 



OPERATING REGION j 



i 





H-OW FRICTpN REGIO^ 



1000 



2000 



3000 4000 5000 
SHAFT SPEED (rad/s) 



6000 



7000 



eooo 



Fig. 3-45 Operating Map of Seal Set No. 4 



^Tliii I •jiiiniiiMWTf'Tiri'rr'-"- I -"^-^""^tiwniV'' ^i'*"*^-^- 



..auAu 



'^' ■' ""^- 




PRltSSUFtE (dbs) 



pogr-TisT 



i 






T 



PERIOD* 






, ! ,■ > |i » ii» |B^»4-* 



U 1. 1. 1 i I ' 



....I,. 



^ 



) 



i i 



10 30 30 40 50 60 70 80 90 100 



TIME (mtn) 
Fig. 3-46 Pressure History - Seal Set No. 4 



r \ 



o 
a 

L 



a. 
in 



7000 



6000- 



5000 



4000 



3000 - 



z 
in 



2000- 



1000 






j 
1 

i 

1 

i 

; 
; 

; 

i 


J 












— — 






j 


7 


•— — 

[ 

! 




■ 


■•*"""• 














: i i 

- ! 1 
ii i . J 



















5 

I 

3 



















•■ 't ! * ■* 
















PRE 






ERI 


3D- 




^ 


Rl 


IN* 


u 




POS 


T-T 


■ST 


PER 


lOD- 




^ 




• ■ 1 

I 1 


- 1 t 1 M 






j 








..... 


i 


1 

1 























4 




— 






















=*• 


->»c 


tsss 


SM 




A 






u 


, , 






.-^ 


l«-»-l 


1 > 


1 ■ t 





-100 -80 -60 -40 -20 10 20 30 40 50 60 70 80 90 100 

TIME (mfn) 

Fig. 3-47 Speed History - Seal Set No. 4 

3-82 



op OQ 

120-, 49-1 



I 

00 
CO 



100- 



80- 



0) 



3 




•4-^ 




(0 




X- 




<U 




Q. 


fin 


t 




0) 




J- 




ca 




0) 




CO 


40 



20- 







38- 



27 



16- 



-7- 



-18.-J 




rad/s rpm 
r-6282 1-60.000 



-5235 



4188 



-3141 



-2094 



-50.000 



-40.000 



30.000 1 

Q. 
CO 

o 
-20,000 ^ 



-1047 -10.000 



Lq 



Time 



Fig. 3—48 Failure of No. 4 Inboard Seal Ring 



S51146 



I 

00 



.002 
.0019t 
.0018 
.0017t 
.0016 
.0015 
.0014 - 
.0013 - 
_.0G12- 
^.OOlli 

^ .ooit- 

O.0009 
_i 

^.0008- 
.0007 
.0006 

.OOOSf 
.0004 
.0003h 
.0002 
.0001 




SEAL SET NO. 4 






f » T - 



-iPRE!-TEST RERIOD 



f*RUN-*H-r 



POST-TEST iPEreiODf- 



— 1 



i: 



I- 
i: 

-Tl- 



■ : 

i: 

i: 

-iJ- 



> * J" J 






iOUTdOARd SEAL 




-r 



-100 



-20 10 20 30 40 50 60 70 80 90 IQO 

TIME (rofn) 



Fig. 3-49 Flow History - Seal Set No. 4 



-i f- 



I 

CO 



CJ 



<x 

I- 

<; 

Of 
UJ 

a. 

UJ 



u 

u. 
a 



UJ 

to 




-20 10 20 30 40 50 60 70 80 90 100 

TIME (win) 



Fig. 3-50 Seal Temperature History - Seal Set No. 4 



I/I 

U 



X 
, J 



a 




-20 10 20 30 40 SO 60 70 80 90 

TIME (mtn) 



Fig. 3-51 Outboard Seal Film Thickness History - Seal Set No. 4 



t— I 
(X 



u 



< 




-100 -80 -60 -40 -20 10 20 30 40 50 60 70 80 90 100 

TIME (mfn) 

Fig. 3-52 Outboard Seal Eccentricity History - Seal Set No. 4 

3-86 



t 

00 

-J 



1/1 
O) 
-Si: 

u 



01 

en 

(O 

S- 
0) 

> 
<: 



.02 
.019 
.018 
.017 
.016 
.015 
.014 
.013 
.012 
.Oil 

.01 
.009 - 
.008 
.007 



SEAL SET NO. 4 
OUTBOARD SEAL 






1- 

I 
f_ 

_._.4.. 






X 

— I- 



-■f-"— 



PREDICIEO CHANGE BASEOj 
ON CENTRIFbGAL GROWTH | 







I 

— i. — 
t 
I 
I 



VERHCM. DIRECT^cm 



1000 



2000 



3000 4000 5000 

SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Kiij. J-53 Outboard Seal Film Thickness versus Hhai t Speed - Seal Set Na. 4 



_J 000« . 




Otb 

Scut 



Runner 



006 
mm 



Inb 
Seal 



Fig. 3-54 

S«al and Runner Motion 
Seal Set No 2 at 3665 rad/s 

• 655 kPa Supply Pressure 

• 103 kPa Otb Drain Pressure 




Otb 
Seal 



Runner 



Inti 
Seal 



Fig. 3-55 

Seal and Runner Motion 
Seal Set No 2 at 4186 rad/s 

• 862 kPa Supply Pressure 

• 517 kPa Otb Oram Pressure 



l; 




inb 
Seal 



Runner 



Otb 
Seal 



Fig. 3-56 

Seal and Runner Motion 
Seal Set No 2 at 47 1 2 raa s 

• 938 kPa Supply Press, e 

• 517 kPa Otb Dram Pressure 



3-88 



ORIGINAL PAGE IS 
OF POOR QUALITY 



ORIGINAL PAGf- rs 

OE POOR yuAu ry 



^^b^^Smmmh ^ ^mmn 


1 



Otb 

Seal 



Runner 



inb 
Seal 



Fig. 3-57 

Sfjai and Runner Motion 
Seal Set No 3 at 3665 rad/s 
e 655 kPa Supply Pressure 
• 517 kPa Otb D' n Pressure 




Otb 
Seal 



Runner 



Inb 
Seal 



Fig. 3-58 

Seal and Runner Motion 
Seal Set No 3 at 3665 rad/s 
o 1206 kPa Supply P'essurp 
• 517 kPa Otb Dram Pressure 



3-89 



ORIGINAL PAGE IS 
OF POOR QU/MTY 



LJ 



L 



MlBB mMI *MM 

■■BnnH 



Oth 
S«ai 



Runner 



Inb 
Seal 




Fig. 3-59 

Seal and Runner Motion 
Seal Set No 3 at 3665 rad/s 

• 1206 i<Pa Supply Pressure 

• 310 kPa Otb Dram Pressure 



Ofb 
Seal 


Fig. 3-60 




Seal and Runner Mono.i 




Seal Set No 3 at 3665 -ac s 


Gunner 


• 1206 kPa SuDpiy ^-bss .<? 




• '03 kPa OlD Oram P-eSS .rt. 


Inb 
Seal 





Li 

Li 

ii 

Li 

li 
fi 



._i 



3-90 



LJ 

u 

» « 

LJ 

Li 



ORrClINAL PACR r«7 
OF PCXJR gUALlTV 




Otb 
Smi 



Runner 



Ino 
Smi 



Fig. 3-61 

Seal and Runner Motion 
Seal Set No. 3 at 4188 -ad. s 

• 1069 kPa SoDOiy P'-«ssure 

• 103 kPa Ott3 Oram ='9ssur9 




GttJ 
Seal 



Fig. 3-62 

Seal anc ^unner Monon 
Seal Set No. 3 ar 4''2 'ic s 

• '■182 <Pa Si-coiv ='9ssLre 

• 51' <P3 Ctt3 "ram ='e93L,r9 



3-91 



_J CC6 L_ 
Aiis I mm I 



d:c6 




ORIGINAL PACE IS 
OF. POOR QUALITY 



Runner 

Vertical 
Probes 

Horizontal 
Probes 



Axis 



Fig. 3-63 

Seal arid Runner Motion 
Seal Set No 4 at 3665 rad s 

• 827 kfa Supply P'essure 

• 517 kPa Oib Dram Pressure 



i I' 



Y Axis 



L. 



■^^n^^Q^^v 



Runner 



Vertical 
Probes 

Ho. izontal 
Probes 



X Axis 



Fig. 3-64 

Seal and Runner Motion 
Seal Set No. 4 at 4188 rad/ s 

• 827' kPa SuDPiy Pressure 

• 517 kPa Olb Oram o-essure 



u 



V Axis 








!■■■ 
!■■■ 

mmK 

!■■■ 


■9 

wm 
■a 


van 

fKMM 

warn 

warn 


■ 

1 




Runner 



Vertical 
Proces 

Horizontal 
Probes 



X Axis 



Fig. 3-65 

Seal irri o.^Tor •.i^*";'- 
Seai Set No 4 .jt -17 1£ -aa i 

• 827 kPa Supply Pres3v.ie 

• 517 kPa Otb Dram P'essu'e 



3-92 



ORIGINAL VAcp r. 



y-A«is 

Hot.zonlai 

P'obes 




Runner 



Vertical 
Probes 

(Near) 

Horizontal 
Probes 

(fan 

X Axis 



Fig. 3-66 

Seal and Runner Motion 
Seal Set No 4 at 5235 rad/s 

• 827 kPa Supply Pressure 

• 517 kPa Otb Dram Pressure 



V-AxiS 

Horizontal 
Probes 



Y-AxiS 
Vertical 
Probes 




Runner 



Vertical 
Probes 

iNean 

Horizontal 
Probes 

(Far) 



Fi^ 



i-b7 



Seal and Runner Motion 
Seal Set No 4 at 5759 rad/ s 

• 724 kPa Supply Pressure 

• 517 kPa Otb Drain Pressure 



3-93 



Seal Ring 




Embedded 

Capacitance 

Probes 



To Oscilloscope 



L; 



Seal Runner 



To Oscilloscope 

Y-Axis X-Axis 

Input Input 



Relative Orbit between 
Runner and Seal Ring 




Oscilloscope Display 
(Identical X and Y Sensitivities) 



Fi>j. i-h6 limbedded Probe Oscillobcope DLsplav Format 



351 '-14 



3-94 



ORICrNAC PAGE IS 
OF POOR QUALITY 



006 







Horizontal 
Z - 20 



Z - 22 



Vertical 
Z - 21 



Z - 23 



Seal Probe Time Trace 
Seal Set No. 4 at 5759 rad/s 

• 724 kPa Supply Pressure 

• 517 kPa Otb Dram Pressure 



Fig. 3-69 



3-95 



C3 



I 




rad/s rpm 
r 5235 r 50.000 



-4188 



-3141 



-2094 



-1047 







Time 



-^1.0 h- 
s 



-40.000 



-30.000 ^ 

o 

CO 

20.000 S 

0) 



-10.000 







Fi>;. i-/il .S|ui-«l- I iiiu- Ciiivc fni" Ty|>ii.il Ai:ce I uT.il ion Kiiii 



BSII49 






[^ i — . : r-^ 



cr 



DRJOINAL PAGE IS 
DF IHX)R gUAUTY 




Fig. 3-71 Damaged Seal; Oucboard Ring - Seal S«C No. 1 




Fig. 3-72 Undamaged Seal; Inboard Ring - SeaL Sec No. 



3-97 





Pig. 3-73 OuMged Runner; Inboard Left, OuCbo«rd Right 
Sc«L Sec No. 1 




Fig. 3-74 Magnified View (X11.4) of Damaged Ranner 
Seal Sec No. I 



3-98 



r' 



•~ t • — •• 



CO 

I 

vC 
O 



iS£Si#:s 




82 

> 

r 



n 
n 



3 



Fig. 3-75 Undamaged Seal; Outboard Ring - Seal Set No. 2 




L- 



a ^ 



Pig. 3-76 DAiMged ScaL: Outboard Ring - S««l Se: No. 3 



Ll 




Fig. 3-77 Undamaged Seal; Inboard Ring - Seal Sec No. 3 



Ll 



II 



li 
h 



3-100 



ORrCTNAL PACE 15 
OF P(j()K giAUTY 





Fig. 3-78 Damaged Nc. 2 Runner; Inboard Left, Oucboard 
RighC - Seal Sec No. 3 




Fig. 3-79 Magnified View (Xil.5) of Damaged No. 2 Runner 



3-101 



u 




ii 



L 



Fig. 3-80 Damaged Seal; Inboard Ring - Seal Set No. 4 



t ( 







1 \ 



Fig. 3-81 Damaged No. 3 Runner; Inboard Left, Oucboard Right 
Seal Set No. 4 



3-102 



ORIGINAL PAGE fS 
QE POOR QUALITY 



11 

u 

I 

L 
L: 









— — — Theory (Constant Cla«ranc»») 




_*- 


1 


.--' 


I 


^-' — 




^- —- 












^^' \ '^^^^^ ^ ^* 




y^^^^^^Z^l. 




>^V'''^^^*'**iZ^— — "" "" 




'^ ,i^^** SpiH»d Increasing 



PrestL Drop 



(a) 



& 



<^ 



No 4 Seals 




No 1. 2. and 3 Seals 



Constant Speed 



Pressure Drop 



(c) 




Pressure Drop 



(e) 



— — Actual (Constant Sfseed) 
~ -•'— Theory (Constant Clearance) 



^ •» ^ — ^ — — — — 

^ ^ ^^ \^ — — - 



Speed^ Increasing 



m 4^ 



E 



CO 



It 



0) 

(P 
CO 



Pressure Drop 



(b) 



No. 3 Seals 




Speed Increasing 



Pressure Drop 



Total Flow 
(Both Sealb) 



Otb Seal 
Flow ■ 




^ 



M 



Supply Pressure 



(f) 



Fig. 3-82 Parametric Variations as a Function of Pressure 
Drop anci Supply Pressure 



351591 



3-103 



4.0 ANALYSIS AND DESIGN OF RAYLEIGH-STEP, HELIUM BUFFER SEALS 

4.1 Operating Condiulona 

The principles of operation and general geometry were presented in Section 
2.0. Additional details concerning performance and theoretical development 
are included in References [2], [6], and [?]. 

Refer to Table 1-1 for the geometric and operational parameters. The surface 
speed of 183 m/s (600 fps) and the buffer fluid pressure of 1379 kPa (200 psia) 
are two conditions that extend the state of the art. 

The properties of helium as a function of pressure and temperature pre indi- 
cated on Table 4-1*. The analysis used the helium viscosity «nd density at 
37.8°C (100°F) and at pressures of 1379, 689, and 344 kPa (200, 100, and 50 
psia), with ambient (101 kPa/14.7 psia) downstream pressures. As indicated on 
Table 4-1, helium gas viscosity varies only ±6Z from the value of 37.8 C 
(100°F) and is independent of pressure. Also the density varies only ±10% 
from the value" of 37.8°C (lOO^F) and is proportional to pressure. 

4.2 Design Considerations 

When designing hydrodynamic gas-lubricated floating-ring seals, there are 
several important considerations. Firct, the hydrodynamic forces that are 
generated when the rings go eccentric with the shaft must be sufficient to 
overcome the friction forces between the rings and the stationary walls. This 
is necessary to maintain the rings concentric with the shaft. 

To minimize leakage, the operating film thicknesses must be relatively small. 
Small film thicknesses are also necessary to provide high fluid-film stiffness 
which is desirable for causing the rings to follow shaft excursions. 



♦Tables are presented consecutively, beginning on page 4-17, 



4-1 



The clearances involved are generally smaller Chan centrifugal growth and 
thermal variations in clearance. Therefore, it is important to account for 
both centrifugal growth and thermal distortions in designing the seal. 

Finally, the seal rings must respond to shaft excursions and runouts without 
contact between the shaft and th-^ rings. This requires that dynamic analyses 
be conducted and dynamic response be carefully investigated. 

The analytical process that was used -.n designing the seals was to: 

1. Optimize the Rayleigh-step geometry principally on the basij of 
fluid-film stiffness using steady-state fluid-film gas bearing theo- 
ry. 

2. Include the efforts of centrifugal and thermal variations in shaft 
and ring geometries in establishing predicted performance. These 
effects change the operating film thickness which is a sensitive 
parameter with respect to performance. 

3. Conduct dynamic response analyses to ensure chat the rings follow 
shaft excursions withaut contact. 

A constraint imposed upon the design of these type seals is to keep the axial 
length as small as possible to minimize shs?t length and not compromise rotor- 
dynamics of the shaft bearing system. In considering both the 50- and 20-miTi 
seals, it was necessary to maintain the axial length of each ring within 
approximately 12.7 ""ti (0.5 in.). This was considered practical number for 
incorporation into actual pump m/ ^hinery. 

4.3 Analys i s and Desig n of the 50- mm Floatin g-R ing Helium Purge Seal 

4.3.1 Rayleigh Step Optim i zat ^ . Studie s 

Initially, studies wer"^ made to determine the optimum Rayleigh-step geometry. 
The parameter optimized was the fluid-film stiffness, which is most important 



4-2 



to prevent ah*ft/ring contact. Figure 4-1* shows dimensionless stiffness as a 
function of the step clearance /I and clearance ratio. The step clearance 
includes the step height plus the film thickness above the step. The curve 
indicates that the step clearance should be 2.65 x the land clearance. The 
seal was designed for an operating land film thickness of 0.0127 mm (0.0003 
in.), so the step height should have a nominal value of 0.0210 mm. The stiff- 
ness values are fairly symmetrical about the optimum point so that the toler- 
ance range could be on either side of the optimum value. The actual step 
height selected was 0.0229 to 0.0254 mm (0.0009 to 0.001 in.), to allow for 
variations in actual operating clearances. These step dimensions provide good 
stiffness over a wide clearance range. 

Figure 4-2 shows the effects of the axial length ratio which is the step 
length/pad length ratio = Ll/L, as a function of dimensionless stiffness. 
This parameter is quite insensitive, although the optimum length ratio is 
0.66. 

Figure 4-3 shows the effect of the circumferential length of the Rayleigh 
step. Over the range shown, it is an insensitive parameter, but the optimum 
ratio of step length/pad length is 0.82, which indicates that long steps are 
desirable. 

The step dimensions were optimized for concentric bearing stiffness. The 
dimensionless length, width, i^nd depth of the step have the following optimum 
values: 

• Depth of step, step clearance/land clearance = 2.65 

• Axial extent of step, Latep/^pad ~ 0.b6 

• Circumferential extent of step, 9 ^ /9 , = 0.82 

*^' Ttep pad 

Studies were also made varying the total numbers of pads. Increasing the 
number of pads from 4 to 5 (while maintaining the same ratio of groove width to 
circumferential pad extent and the same step geometry) decreased the centered 



*Figures are presented consecutively, beginning on page 4-22, 



4-3 



stiffness by only IX, indicating that four pads yield fairly optimum fluid 
film performance. The four-pad geometry was selected. Figures 2-1 and 2-2 
indicate the nominal dimensions used. 



A. 3. 2 G eneral Confi guration and Design 

Figure 4-4 indicates a developed view of the inner surface, including the 
hydrodynamic geometry and also shows the end wall contact surface region in 
larger scale. The contact surface was maintained as small as practical (0.762 
mm) (0.030 in.) and as close to the shaft as possible to reduce the maximum 
thrust loading on the seal rings. 

Figure 4-5 shows the installation of the rings in the seal tester. What is to 
bn particularly noted is the diameter of the seal runner which is considarably 
larger than the shaft diameter used in the experimental rig. The stiaft size 
is approximately 30 mm (1.18 in.) in diameter while the runner size is 
required to be 50 mm (1.97 in.) in diameter. Examination of this figure indi- 
cates that centrifugal growth of the seal runner will be significant. The 
seal runner configuration was designed to provide equal distribution of the 
centrifugal growth without causing closure at the ends of the runner. 

Figure 4-6 indicates what centrifugal growths will do to the runner at the 
maximum operating speed of 7330 rad/s (70,000 r/min). It will expand radial- 
ly, approximately 11.43 Mm, with slight variations at the end (1.88 Mm). The 
end closures are inconsequential and confirm the advantage of the taper, T, 
configuration employed. The total radial expansion, however, is of the same 
order of magnitude as the film thickness, thus manufacturing and installation 
dimensions must be large enough to accommodate the runner growth or clearance 
closure. The runner was shrunk over a flexure which could compensate for 
increase in the inside diameter of the runner due to centrifugal expansion. 

The dimensions of shaft and ring indicated on Figure 4-5 were selected to 
provide near optimum clearance at the operating condition considering the 
effects of both thermal contraction and centrifugal growth, respectively. 



4-4 



The seal ringa were maUe entirely of carbon (i.e., no metallic bands) to 
enhance dynamic response characteristics (Pure Carbon P5N). The runner was 
made of InconeL 718 for strength purposes and coated with a layer of tungsten 
carbide with a chromium binder. The ring mating housings were Inconel 600. 

The final seal detail drawings were completed by the seal manufacturer, Stein 
Seal Company of Philadelphia, Pa. Figure 4-7 shows the assembly drawing of 
the 'jeal and housings. Figure 4-8 shows a detail drawing of the outboard seal 
ring. A photograph of one set of seal rings was previously shown on Figure 
2-3. 



4.3.3 Fluid-Film Performance 

Fluid-film performance was initially established using a nominal radial 
clearance of 12.7 Mm and correcting for centrifugal growth of the shaft 
sleeve. Subsequently, larger clearances were examined to allow for variations 
in clearance due tn tolerances and to thermal contractions of the sleeve. 
Figure 4-9 shows the variation in concentric seal clearance due to centrifugal 
expansion of the sleeve as a function of shaft speed, presuming a 12.7 Mm 
operating clearance at the maximum speed (7330 rad/s) (70,000 r/min) condi- 
tion. The assembled radial clearance is approximately 29 Mm to attain a 12.7 
Mm clearance at operating speed. This assembled clearance does not account 
for thermal contractions. It was subsequently found that the installed clear- 
ance should be as small as possible, to achieve an operating clearance of 
approximately 12.7 Mm, because thermal contractions of the runner had a stron- 
ger influence than centrifugal growth. 

Figure 4-10 shows the fluid-film force developed in a seal ring versus eccen- 
tricity ratio at three different helium pressure levels, 344, 689, and 1379 
kPa (50, 100, and 200 psia). Superimposed on these curves, are frictional 
resistance forces between the floating rings and the stationary housing. The 
radial clearar was 12.7 Mm. The resulns indicate that maximum frictional 
resistance can be overcome by hydrodynamic forces at high-speed operation. It 
also indicates that at 1379 kPa (200 psia) buffer pressure, low-speed hydrody- 
n£;iiic forces will not overcome frictional resistance. At maximum speed of 



4-5 



7330 rad/,9 (70,000 r/min) and maximum butter pressure at 1379 k.Pa (200 pfsia), 
an eccentricity ratio of approximately 0.65 is necessary to overcome the maxi- 
mum Erictional resistance of approximately 42.3 N (9.5 lb). Table 4-2 shows 
the effects of radial clearance on the seal's ability to overcome frictional 
resistance . 

The tabulated values of eccentricity ratio and minimum film thickness are 
necessary to produce sufficient hydrodynamic forces to overcome the maximum 
friction forces at each of the buffer fluid pressures indicated. The results 
clearly indicate Che superior performance at the low clearance condition of 
12.7 Urn. At the high buffer fluid pressures, the low clearance installation 
produces the higher minimum film thickness. At a 25.4 ym radial clearance, 
the seal would not adequately overcome frictional resistance at a buffer pres- 
sure of 1379 kPa (200 psia). 

Viscous power loss for a single ring as a function of speed and pressure is 
indicated on Figure 4-11. At maximum speed and pressure the total power loss 
is approximately 85 W. 

Seal leakage, on the basis of laminar flow without inertia drop losses, is 
shown on Figure 4-12. A reduction in leakage with speed occurs because the 
centrifugal runner growth causes closure of the clearance as the speed 
increases. At 7330 rad/s (70,000 r/min) with an operating clearance of 12.7 
\lm, the seal leakage (single ring) is 0.908 x lO"-* kg/s (11.5 scfm). The 
leakage values on Figure 4-12 assume concentric operating clearance (i.e., the 
rings are cpnrered with respect to the shaft). The leakage increases approxi- 
mately 40Z from concentric to full eccentric operation. 

Testing indicated that seal leakage was lower than prediced. Modifications 
we.-e made to the analysis to includ inertia effec , at the seal dam inlet and 
in the film itself. The analysis to accomplish this is described in Appendix 
A. Correlation between experiment and theory was then significantly improved. 

Figure 4-13 shows fluid temperature rise as a function of speed and pressure. 
These curves assume all heat generated by viscous friction is absorbed by the 
flow leaking through the clearance annulus; heat transferred to the shaft is 



4-6 



.J 



not accounted for. Therefore, the temperature ri.'i«'3 indicated on Figure 4-13 
are exaggerated. Since flow increases with buffer pr«;jfjure, the temperature 
rise is an inverse function of buffer pressure and will be lower as the pres- 
sure increases. The results indicate that high-speed, low-pressure operation 
IS to be avoided. Table 4-3 indicates maximum operating speed as a function 
of buffer pressure to limit the temperature rise to 22°C (39.6°F), which can 
bo considered a maximum safe value. The values of flow used in the temper- 
ature rise computations were based on viscous laminar theory without inertia. 
When inertia effects are included, there is a flow reduction which would 
result in a proportional increase in temperature rise. This increase in 
temperature rise was one of the contributing factors to clearance closure that 
was experienced by the inboard ring of seal No. 4. 

4.3.4 Thermal Analy sis 

The model for the thermal analysis is shown on Figure 4-14. The four signif- 
icant modes are 6 and 7 for the two rings and 9 and 10 for the runner. Rela- 
tively hot helium enters the space between the two rings and flows inboard and 
outboard through the annular clearance between the rings ?nd rotating collar. 
On the inboard side near node 9, the collar is cooled by the LOX escaping from 
the bearing, while on the outboard side near node 10, helium at much higher 
temperature escapes. The analysis considered 0.0018 kg/s (22.74 scfm) of 
helium entered the buffer chamber at an inlet temperature of 21°C (69.8°F) and 
that approximately 90 W of heat was generated at the fluid-film interface. 
Studies were made to establish the sensitivity of varying parameters. Table 
4-4 lists the cases that were run on the computer and the results obtained. 

The table indicates that the resulting temperatures are not very sensitive Co 
the values of the heat transfer coefficients used. 

Seal operating film thicknesses have been calculated using various values of 
installed clearance and accounting for centrifugal growth and thermal 
effects. Recommended dimensions are indicated on Table 4-5. The following 
nomenclature applies: 



4-7 



^^ring ~ ID "f the .'jeal ririj^ 
Dghaft ~ ^D ' the shaft runner 
Cq = Asaembled radial cLnarance 

Gq' = Radial clearance including centrifugal growth of shaft 
C()" = Running radial clearance including centrifugal and 
thermal growth 

Each set of three consecutive lines in the table corresponds to the maximum, 
average, and minimum clearance of the tolerance range. These calculations of 
running clearance used the following data: 

• Radial growth of shaft runner at 7330 rad/s (70,000 r/min): 
11. A3 Um 

• Mean coefficient of thermal expansion of carbon graphite: 
5.58 X 10~^/°C (3.1 X 10~^/°F) 

• Mean coefficient of thermal expansion of Inconel 718: 
12.78 X 10"^/°C (7.1 X 10~^/°F) at 21°C (69.8°F) 
10.62 X iO~-/°C (3.9 x 10"°/°F) at -196°C (-320. 8°F) 

The difference in operating clearances on Table 4-5 are due to differences in 
nominal ring and shaft dimensions. The middle line corresponds to nominal 
size with tolerance variation indicated above and below. Table 4-5 also indi- 
cates recommended final dimensions to account for both centrifugal and thermal 
growths . 

The results clearly indicate that the assembly clearance be 5 to 8 )Jm (as 
small as possible for assembly without requiring an interference fit). That 
is, the dimension of the ring ID should be 50.02022/50.0253 mm (1.9693/1.9695 
in.), while that of the shaft runner OD should be 50.00598/50.01006 mm 
(1.9693/1.9689 in.). The running seal clearances are then reduced to 11.176 
and 25.4 Um on the air and LOX sides, respectively. In the final design, it 
was necessary to compromise further for manufacturing purposes. The outboard 
seal ring was designed for higher clearance than the inboard seal ring; since 
thermal effects would cause increased clearance on the inboard ring. The 
final clearances employed are discussed in Section 2.0. 



4-8 



4^3.5 Seal Ring Dynamic Uoriponfii; 

Because ot high fiurtacc upeeJa and low operatinji; tilm t.hicknerjfj , it wa^i impor- 
tant to investigate dynamic response and to design a system that avoids 
rubbing contact. A time-transient uiaLysis was employed, whereby time is 
discretized into small increments. At each increment fluid-tilm and friction 
forces aru examined to determine net forces on the ring. These forces can 
then be inserted into the equations of mut ion to establish ring motion for 
that time increment. Thus, a time history of ring movement is provided as a 
function of shaft excursions. If fractional forces exceed fluid-film force?; 
at any instant of time, ring motion is curtailed. 

Studies indicated that for any particular eccentricity ratio, the fluid-film 
force did not vary strongly with angular position and using a constint valu(> 
would produce accurate results. This was done in the computer analysis, but 
the conservative or the lowest values of film forces were used for each eccen- 
tricity ratio. Figure 4-15 shows typical fluid-film force curves that were 
applied in the program. Shown are the radial component (along the line of 
centers) and the tangential component (normal to the line of centers) as i 
function of eccentricity ratio at the 1379 kPa (200 psia), 7330 rad/s (70,000 
r/min) operating condition. Note the tangential component is small. 

Force information was interpolated in the computer code, so that knowing the 
shaft location produced the normal and tangential forces required. Similar 
information was produced at other pressures and speeds. 

A shaft orbital radius of 2.b-' Urn simulates anticipated test conditions. 
However, much larger orbital amplitudes were also examined to establish an 
adequate "factor of safety" with respect to dynamic response. 

Figure 4-16 shows a typical orbital response case when the shaft runout is 
7.62 pm, operating at 7330 rad/s (70,000 r/min) and 1379 kPa (200 psia) buffer 
fluid pressure. The starting position was with the seal ring concentric with 
the shaft. The ring settles into a rectangular shaped orbit with a maximum 
eccentricity ratio of 0.625. The eccentricity and orbit indicated are rela- 



4-9 



t ivn to the fihatc, and tin; c Icuiratimr circle, betwt'.en t lit' uiil'iiilt' lirclf ,iii(i 
t hu orbit: iu tho. clearance aeen by an observer riding on t lu; 'jhatt. . 

I'igure 4-17 shown a cage at 345 kPa absolute (50 psia) butter privisurt' at a 
fjhaft; <!peed ot 2094 rad/a (20,000 r/min). At t:he lower speed condition hydro = 
dynamic torces and fjtiftneaa arc relatively low, <jo thai tin; rinj?, can b(! 
expected to respond rather sluggishly to shaft motions. This is demonstrated 
by the case shown on Figure 4»17. In this instance, the shaft radial runout is 
0.0102 mm (0.0004 in.). The starting position was at a shaft eccentricity 
ratio of 0.95. The ring moved slightly more concentric, and then maintained 
its position without further motion. The shaft is moved eccentrically within 
the clearance circle at its prescribed eccentricity ratio of 0.8. Basically, 
at these conditions film stiffness was insufficient to center the ring and 
permit it to follow shaft motions in a nearly concentric position. For this 
case, the 345 kPa absolute (50 psia) .utfer pressure will not produce high 
friction forces', therefore, at higher pressure levels the seal ring will 
continue to remain stationary. 

Figures 4-18 and 4-19 show response at low speed and relatively high pressure 
conditions so that friction forces dominate. Th(» orbital relationship between 
the shaft and ring is erratic and complex with reverse loops involved with 
each cycle. 

Studies were also made of composite rings where an Inconel ring was shrunk 
around the outer periphery ot the carbon ring. This was done to produce more 
uniform thermal distortions between the rings and shaft. The added mass ot 
the composite rings, however, caused contact failure due to excessive inertial 
accelerations that would not permit the rings to move in unison with the 
shaft. A typical case is shown on Figure 4-20. Contact occurs before an orbit 
cm be completed. Thus, it is important to keep the ring massas low as possi- 
ble, and integral carbon rings without metal shrouds are necessary for 
adequate dynamic response. 

Table 4-6 summarizes the operating conditions and results of the dynamic 
response computer runs. The variables included shaft eccentricity or runout, 
operating clearance, buffer fluid pressure and speed. The orbit eccentricity 



4-10 



is the iMxinua eccentricity of the .'esulting steady-stnte orbit of the ring. 
The ainiauM film thickness, h^, was the ffliniiDun value experienced during the 
orbital response. Potential problem areas, where niniimun films are becoming 
dangerously sitall or are negative, are underlined. The 1379 luPa (200 psia), 
7330 rad/s (70,000 r/oin) condition is considered acceptable; it might be a 
narginal problem at very high shaft runouts, which were not anticipated to 
occur. In fact, they did occur and caused failure of Seal Set Ho. 1. 

Figure 4-21 shows the effects of /arying the shaft runout. Note that at a 2094 
rad/s (20,000 r/fflin) and 344 kPa (SO psia) buffer pressure, the eccentricity 
ratio is becoming very high and is an operating condition that should be 
avoided. In this instance, inertia-dominated motions of the ring are overcom- 
ing the friction retardation forces and the fluid-film stiffness capacity. At 
70,000 rad/s (7330 r/min) and 1379 kPa (200 psia), the limit cycle is well 
controlled even with a high shaft eccentricity of 12.7 Urn. An overall suimary 
plot is shown on Figure 4-22. It is a plot of maximuin transient orbital eccen- 
tricity ratin versus shaft vibration or runout orbit for varying types of 
rings (solid carbon or composite carbon and metal) ac different pressure 
conditions. Composite rings are sometimes used to more nearly equalize ther- 
mal expansions between the rings and shaft. The operating speed is 7330 rad/s 
(70,000 r/min). 

Note chat the composite carbon/Inconel rings have significantly less toler- 
ance to shaft orbit than do the solid carbon rings. Note also that higher 
pressure and consequently higher frictional forces are beneficial (at high 
operating speeds) because the friction forces prevent excessive inertial 
response of the rings. Thus, there are really two limiting conditions 
concerned with ring design. First, the fluid-film forces should be great 
enough to overcome frictional resistance to ensure against contact, and 
second, there should be sufficient friction to prevent inertia dominated 
motion of the ring that could cause contact under high-speed conditions. 

All of the dynamic discussion thus far has presumed an operating radial clear- 
ance of 12.7 Urn. Because of the difficulty of obtaining this clearance 
precisely, due to manufacturing tolerances, centrifugal . growth and thermal 
contractions, studies were also made at varying clearances. The dominating 



4-11 



influence is the thermal effects which causes a reduction in shaft diameter 
and an opening of the clearance. Thus, several computer runs were made at 
2S<4 Mm radial clearance or twice the designed clearance. 

The principal result was that the rings are forced into a concentric position 
without contact, and they remain fairly stationary in that position with shaft 
orbits inside the clearance volume. Operation at larger clearances is safe 
from a dynamic standpoint. 

4.3.6 Summary of Results and Conclusions of Analytical Studies 

Performance at a design clearance of 12.7 Um is very good and can satisfy all 
requirements. Also, performance will be satisfactory over the tolerance range 
specified. The two limitations on performance are as follows: 

1. Insufficient hydrodynamic forces to overcome friction forces - a 
low-speed, high-pressure constraint 

2. Insufficient f ri( cion to counteract inertia forces - a high-speed, 
low-pressure constraint. 

Figure 2-4 showed an operating range map seal that accounts for aJ 1 
constraints. If the pressure follows a speed squared relationship to a maxi- 
mum of 1379 kPa (200 psia) at 7330 rad/s (70,000 r/min), it is expected that 
seal performance will be satisfactory. 

Centrifugal runner growth and thermal contractions significantly affect oper- 
ating clearances and they must be considered in performing the analysis. 
Centrifugal forces on the runner tend to close the clearances while thermal 
contractions tend to open the clearances. Thermal effects have the stronger 
influence, requiring installation clearances to be as small as practical (10.2 
- 22.9 Vim diametral clearance). 

Low mass rings are necessary to dynamically track runner excursions. The 
rings should be made of carbon without composite metallic rings on the outer 
circumference. 



4-12 



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4.4 Ana lysis and D esign of t he 2 0-mm Float in);-Kin >> Helium {'urga Seal 

4 .4. 1 Genera l Conf i guration and Op e r a t ing C o nditions 

The general configuration of the 20-mni helium purge seal is shown on Figure 
4-23. A separate shaft would have been used for testing of the 20-mni design 
and the buffer rings were Intended to mate directly against the shaft. In 
other respects, the design is very similar to the 50-mm with a porportional 
size reduction. The optimization parameters were the same as for the 50-miii 
design. 



a The 20-mm seal was constrained in surface speed by limitrtions on the maximum 
operating s^j^ed of the test rig. The maximum design speed of the rig is 10,472 
rad/s (100,000 r/min). A 20-mni shaft rotating at 10,472 rad/s (100,000 r/min) 
will produce a surface speed of 105 m/s (344 ft/s) which is only 57% of the 183 

_ m/s (600 ft/s) specified. Limitations on surface speed also limits hydrody- 

I namic force generation, which in turn limits the buffer fluid pressure or 
allowable friction force at the contact interface. It appears that the abso- 

■ lute maximum pressure would be 6£,'),5 kPa (100 psia) and the allowable pressure 
will further reduce as speed decreases. 

Another factor concerned with the 20-mni seal is centrifugal growth of the 
runner will be negligible. This is due to the runner of the 20-mm seal being 
integral with the shaft, and because of limitations on the maximum shaft 
surface speed. 

Although Che 20-nim seal is significantly smaller than the 50-mm seal, the 
nominal operating clearance of 0.0127 mm (0.0005 in.) remains the same, 
because this was considered to be the smallest practical value for safe opera- 
tion. 

4.4.2 Fluid- F ilm Performance 

Figure 4-24 shows the fluid-film forces as a function of the eccentricity 
ratio and operating speed. The effects of buffer pressure arc Indicated, but 



4-13 



variations in hydrodynamic flaid~film force due to this parameter are practi- 
cally insignificant. The maximum spevid examined was 10,472 rad/s (100,000 
r/min), which is the limit of the test rig. Superimposed upon this curve are 
Che contact fric'.ion forces at 689.5 kPa (100 psia) and 344.7 kPa (50 psia) 
buffer fluid pressure levels. At a buffer pressure of 1379 kPa (200 psia), 
there is insufficient hydrodynamic capability to overcome the contact fric- 
tion force. Maximum load capacity at 10,472 rad/s (100,000 r/min) is approxi- 
mately 13.34 N (3 lb), which occurs ac an eccentricity ratio of slightly over 
0.9. To provide a reasonable safety margin the maximum buffer was designed to 
be 689.5 kPa (100 psia) which would require an eccentricity ratio slightly 
below 0.8, or a minimum film thickness of approximately 0.0025 mm (0.1 mils) 
to move the ring into a concentric position. 

Figure 4-25 indicates what occurs if the buffer fluid pressure is increased as 
the square of the speed, the anticipated method of bringing the test rig up to 
speed. Two pressure-speed curves are shown, one for a 345 kPa (50 psia) pres- 
sure at 10,472 rad/i (100,000 r/min) anii one for a 689 kPa (100 psia) pressure 
at 10,472 rad/s (100,000 r/min). For the 345 kPa (50 psia) situation, the 
eccentricity ratio remains safely between 0.5 and C.,5. For the 689 kPa (100 
psia) case, the eccentricity ratio will exceed 0.75 at approximately 6597 
rari/s (63,000 r/min), and as indicated on Figure 4-25 will approach 0.8 before 
it overcomes the anticipated friction force. This will probably be an accept- 
able situation, because the ring moves to a concentric position after the 
iriction force is )/ercome. 

Figures 4-26 and 4-27 show the effects of clearance variations on ring load 
capacity. Figure 4-26 is tor a constant operating speed of 10,472 rad/s 
(100,000 r/min). Note the significant decrease in load capability as the 
clearance is increased. At a clearance of 0.0191 mm (0.00075 in.), the maxi- 
mum allowable pressure is more like 345 kPa (50 psia) rather than 689 kPa (100 
psia). At a clearance of 0.0254 mm (O.'^Ol in.), a maximum pressure of 206.8 
kPa (30 psia) appears to be appropriate. Figure 4-27 indicates fluid film 
load capacity as a function of speed and eccentricity at a clearance of 0.0254 
mm (1 mil). This curve demonstrates that it is necessary to operate at low 
pressure levels to overcome friction at high clearance conditions. 



4-14 



1 



Anticipated leakage flow through one ring is shown on Figure 4-28. This curve | 
applies •• an eccentricity ratio of 0.5. At the helium pressure levels that 

the rig will operate at 689 kPa (100 psia), the leakage is only 0.000118 kg/s p 

(0.00026 Ib/s). At a maximum pressure level of 889.6 kPa (200 psia), the i 

leakage is 0.000363 kg/s (0.0008 Ib/s). If the clearance is double, the flow I 

r 

will increase by a factor of 8, so that it will increase to approximately ? 

^' 

0.000908 kg/s (0.002 Ib/s) at 689 kPa (100 psia). This assumes laminar i 

viscous flow. The actual flow will be less when account is taken of inertia ^ 

effects at entrance and in the film, as was done for the case of the 50-mm 

seal. Leakage as a function of eccentricity ratio is shown on FiF,ure 4-29. 

From fully eccentric to fully eccentric, the leakage increases by a factor of 

1.7. 

Figure 4-30 shows power loss as a function of speed at an eccentricity ratio '^1 

of 0.5. The power loss is approximately 12 W at 10,472 rad/s (100,000 r/min). " 

Fluid temperature rise is shown on Figure 4-31. As with the 50"-mm design, 

this curve was based on the assumption that all the heat goes into the flowing 

fluid, while in reality a good deal of it will be transferred into the cool 

shaft. High temperatures are predicted for the low-pressure, high-speed .| 

conditions where there is high heat generation anii lov. flow. 4 



4.4.3 Dynamic : esponse 

Rotor dynamic studies indicate maximum shaft amplitudes will probably be lejs 
than 0.00254 mm (0.0001 in.). 

Figure 4-32 shows the radial and tangential force magnitudes of the fluid film 
obta.ned from the steady-state computer code, as a function of eccentricity 
ratio. These values were interpolated for use in the dynamic computer code. 

A spectrum of response computer runs were made over a range of varying pres- 
sures, speeds and ring clearances. They are summarized on Table 4-7. Exam- 
ination of the table indicates acceptable response except at high shaft 
runouts. Eg, .low pressures, Pq, and relatively high operating speeds and at 
high runouts, low speed, and high pressures. Problem situations are under- 



4-15 



lined. At high-speed, low-pressure conditions, ring inertia forces predomi- 
nate over the retarding friction force so that the seal ring runs away around 
I'he shaft and ultimately contacts. These are operating conditions which are 
to be avoided. 

Figures 4-33 and A-34 graphically display the summary results with a shaft 
runout of 0.00254 mm (0.0001 in.). Figure 4-33 shows that the limit cycles 
become Larger as speed reduces and as pressure increases. This indicates a 
degradation of the hydrodynamic capacity to permit low amplitude or concentric 
response with the rotating shaft. la all cases > however, there was sufficient 
capacity to prevent contact between the ring and shaft. Figure 4-34 shows a 
cross-plot with pressure as the abscissa. Again the plot clearly shows the 
higher amplitudes at the lower speeds and higher pressure conditions. 



^~^j 



4-16 



I 



TABLE 4-1 



PROPERTIES OF HELIUM 



Temperature 
(°C) 


Pressure 
(kPa) 

.1379 


Mass Density 
(kg/m^) 

2.34 


Viscosity 
(MPa-s) 


10 


19.4 


10 


689 


1.17 


19.4 


10 


344 


0.58 


19.4 


37.8 


1379 


2.13 


20.7 


37.8 


689 


1.07 


20.7 


37.8 


344 


0.53 


20.7 


65.6 


1379 


1.96 


21.9 


65.6 


689 


0.98 


■ 21.9 


65.6 


344 


0.49 


21.9 






4-17 



• TABLE 4-2 

REQUIRED ECCENTRICITY AND FILM THICKNESS 
TO OVERCOME FRICTIONAL RESISTANCE 





Maximum 


Radial 






Buffer 


Frictional* 


Clearance 


Eccentricity 


Minimum Film 


Pressure 


Force 


Co 


Ratio 


Thickness h^in 


(kPa) 


(N) 


(Mm) 


(e) 


(Um) 


1379 


42.26 


12.7 


0.6 


5.08 


689 


21.13 


12.7 


0.35 


8.26 


344 


10.59 


12.7 


0.17 


10.54 


1379 


42.26 


19.05 


0.87 


2.49 


689 


21.13 


19.05 


0.60 


7.62 


344 


10.59 


19.05 


0.32 


12.95 


1379 


42.26 


25.4 


0.99 


0.254 


689 


21.13 


25.4 


0.76 


6.10 


344 


10.59 


25.4 


0.48 


13.21 



'•'Coefficient of Friction = 0.2 



TABLE 4-3 



MAXIMUM OPERATING SPEED FOR AT = 22 °C 



Buffer Pressure 
(kPa) 


Speed 
(rad/s) 


344 


5,131 


689 


6,283 


1379 


7,330 



4-18 



TABLE 4-4 



RESULTS OF 50-MM THERMAL ANALYSIS 



Case No. 


He Inlet 
Temperature 
(°C) 

21 


6 


Node Tempi 
7 


arature 
9 


(°c) 

10 


Remarks 


1 


-9.8 


-10.3 


-41.4 


-92.1 


Base case 


2 


21 


-9.8 


-10„3 


-40.4 


-90.3 


Recalculated 
He-Fluid veloc- 
ity 60% of 
runner speed. 


3 


21 


-9.8 


-10.3 


-42.6 


-95.1 


Same as Case 2, 
but all he's 
doubled. 



TABLE 4-5 



RECOMM ENDED DIMENSIONS ACCO U MTING FOR 
CENTRIFUGAL GROWTH AND THERMAL "CONTRACTIONS 



Dring 
(um) 


Dshaft 
(um) 


Co 
(Um) 


Tshaft 
(^C) 

Air Side 


"(is? 


Co' 
(Um) 


Co 

(um) 


50.02430 
50.02276 
50.02002 


50.00498 
50.00752 
50.01006 


10.16 
7.62 
5.08 


-41.44 
-41.44 
-41.44 

LOX Side 


-9.83 
-9.83 
-9.83 


-1.27 
-3.81 
-6.35 


13.72 

11.20 

8.66 


50.02530 
50.02276 
50.02002 


50.00498 
50.00752 
50.01006 


10.16 
7.62 
5.08 


-92.11 
-92.11 
"92.11 


-10.28 
-10.28 
-10.28 


-1.27 
-3.81 
-6.35 


27.97 
25.45 
22.91 



4-19 



TAB LE 4-6 
50-MM SEAL TRANSIENT ANALYSIS: SUMMARY OF RESULTS 











Orbit 


Mini mum 




Shaft 


Pressure, 




Eecentrici ry 


Film 


earance, Cq 


1 Eccentricity' 


Po 


Speed, N 


Ratio 


Thickness 


(Jim) 


(Mm) 


(kPa) 


(r-d/s) 


(e) 


(Mm) 


12.7 


2.54 


1379 


2094 


0.63 


4.7 






1379 


3142 


0.64 


4.6 






1379 


3236 


0,26 


9.4 






1379 


7330 


0.24 


9.7 






689 


2094 


0.46 


6.9 






689 


3142 


0,28 


9.1 






689 


5236 


0.25 


9.5 






689 


7330 


0.22 


9.9 






344 


2094 


0.26 


9.4 






344 


3142 


0.24 


9.7 






344 


5236 


0.22 


9.9 






344 


7330 


0.25 


9.5 






172 


2094 


0.22 


9.9 




5 = 08 


344 


2094 


0.49 


6.5 






1379 


7330 


0.40 


7.6 




7.62 


1379 


7330 


0.63 


4.7 



10.16 



12o / 
2.54 10.16 



12.7 



344 


7330 


0.63 


4.7 


344 


2094 


0.62 


4.8 


689 
344 


7330 
2094 


0.43 
0.85 


7. ,2 
1.9 


1379 


7330 


0.65 


4.4 


1379 


7330 


0.66 


4.3 


517 
448 
34^ 


7330 
7330 
7330 


0.44 
0.42 
1.02 


1.4 
1.5 
-0- 


379 


7330 


0.43 


1.4 


344 


2094 


0.85 


3.8 


689 


7330 


0.52 


1.2 


1379 


7330 


0.60 


1.0 


517 


7330 


0.60 


1.0 



482 7330 0.58 1.1 



4-20 



TABLE 4-7 



20-MM SEAL TRANSIENT ANALYSIS: SUMMARY OF RESULTS 



Weight of Ring - 0.0146 kg (0.0322 lb) 
Coefficient of Friction, li = 0.2 







Shaft 








Orbit 


Minimum Film 


Clearance 


Eccentricity, 


Pressure, 


Speed, 


Eccentricity 


Thickness, 


Co 


E.J 


Po 


N 


Ratio 


^M 


(u-m) 


(pm) 
2.54 


(kPa) 
345 


(rad/s) 
10,472 


(e) 
0.231 


(urn) 


12.7 


9.65 










689 


' ' 


0.225 


9.9 










1379 


0.259 


9.4 










345 


7,854 


0.220 


9.9 










689 


'1 


0.204 


10.1 










1379 


0.350 


8.26 










345 


5,236 


0.221 


9.9i 










689 


w 


0.241 


9.65 










1379 


0.459 


6.86 










345 


2,618 


0.238 


9.68 










689 






0.546 


5.77 






\ 


r 


1379 






0./35 


3.35 






7.62 


345 


1 


' 


0.639 


4.57 










345 


5,236 


1.013 


-0- 










689 


i 


0.636 


4.62 










345 


7,854 


1.008 


-0- 










517 






0.615 


4.83 










689 






0.648 


4.47 










345 






1.05 


-0- 










517 






1.006 


-0- 


\ 


r 


\ 


f 


689 






0.635 


4.51 


19.1 


2.54 


1379 






0.4113 


11.2 


1 






345 






0.270 


13.97 


25.4 






345 






0.312 


13.21 






1 


1 


1379 


' 


1 


0.364 


12.20 






7.62 


1379 


10,472 


0.55 


8.64 






5.08 


1379 


i 


0.508 


9.40 






7.62 


345 


7,854 


0.459 


10.4 








345 


5,236 


0.490 


9.65 


' 


f 


' 


r 


345 


2, 


618 


0.588 


7.87 



4-21 



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Eccentricity Ratio 



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20 

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689 kPa 

1379 kPa 



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Speed (rad/s) 



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6000 



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I 




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W (kg) 00282 
ES (mm): 00102 
PO (kPa. abs) 345 
Mu 02 



ECCMAX 
ECCMNl 
ECCMX1 



095 

792964 

848226 



Fig. 4-17 50-mm Ring Orbital Response, Wi kPa Absolute (50 psia), 

2094 rad/s (20,000 r/min), 0.0102 am (O.OOO-^ in.) Shaft 'Runout 



•bl604 



2094 rad/s 



I. 




D (mm) 50 
W(mm): 00282 
ES (mm): 00102 
PO (kPa. abs) 689 
Mu: 02 



ECCMAX 0635413 
ECCMN1 0235992 
ECCMX1 460899 



n&- 



-18 50-nun Kinj; Orbital Response; b«9 kPa Absolute (KtO psla); 

2094 rad/s (20.000 r/min); 0.0102 mm (0.00'% in.) Shaft Runout 



as 1603 



3142 rad/s 







D (mm): 500 
W (kg): 00282 
(mm) 00102 
PO (kPa. abs) 1379 
Mu: 02 



ECCMAX 
ECCMN1 
ECCMXl 



637233 
0130857 
635234 



Fig. 4-19 50-nun Ring Orbital Response; 1)79 kP.i Absolute (200 psia) ; 

3U2 rad/s (JO.OOU r/min) ; 0.0254 mm (. 0.U004 in.) Shaft Runout 



K1605 



h) 



7330 rad's 



i 

*- 



Contact 




CarborVlnconel Ring 
ES(mm): 000635 
PO (kPa): 1279 
Mu: 02 



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Helium Pressure kPa (psia) 
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o < 689 ( 1 00) 
345 (50) 






Friction Force 
689 kPa (100 psi) 




Friction Force 
345 kPa (50 psi) 



10.472 rad/s 
(100.000 rprD) 



7.854 rad/s 
(75.000 rpm) 



5.236 rad/s 
(50.000 rpm) 



2.618 rad/s 
(25.000 rpm) 



1 2 3 4 5 6 7 8 9 1.0 

Eccentricity Ratio 



^'^;■ '•--A 20~wm IK-Hum Soal; Kiir« t- vtr.wiis lu-iAtilri> ity 
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N' ^ Speed (rad/s) 




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Speed (rpm) 



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Speed (rad/s) 



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Eccentricity Ratio 



Fig. 4-26 20-nK! Helium Seal; Effect of Clearance on Fluid Fila Force; 
10.472 rad/s (100.000 r/min); \ i7<* kPa Absolute (200 psia) 






3 5r 

30 

25 



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Eccentricity Ratio 



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20 40 60 80 100 120 140 160 180 200 
Helium Pressure (psia) 



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Helium Pressure (kPa) 



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1400 



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Speed (rad/s) 



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Eccentricity Ratio 



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- 869 kPa (100 psia) 



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Friction Coefficient 0.20 
Seal Ring Weight 0145 kg 
Shaft Orbit Radius 0025 mm 



345 kPa (50 psia) 




J L. 



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20.000 40.000 60.000 
Speed (rpm) 



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'40 60 80 100 120 140 160 180 200 

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Helium Pressure (kPa) 



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a5isa9 



5.0 ANALYSIS AND DESIGN OF SPIRAL-CROOVK LUX SEALS 

5.1 General Discussion 

The spiral-groove face seal is a prime candidate tor application to LOX turbo- 
pumps. It IS a flaid-tilm seal that can ettectively inhibit leakat(e and avoid 
rubbing contact that could cause catastrophic explosion failure in a LOX envi- 
ronment. As described in the introduction, the function of the LOX seal is to 
aid in preventing leakage of LOX from the pump end of the machine. Require- 
ments and operating conditions were presented in Section l.O. 

The most demanding requirements that extend the present state ot the art 
include the high relative surface speed of 183 m/s (600 ft/s) and the high 
pressure of 3.17 MPa (750 psig) to be sealed. 

Originally, MTI examined a conventional type ot spi ral -j^jroove seal that was 
labeled the straight-through de-iij^n. the spiral grooves extended to the 
outside diameter and the fluid was pumped inward to a dam region at the inte- 
rior ID of the seal. Althou^^h excellf^nt pertormancf characteristics were 
predicted, the straight-through designs were ultimately abandoned because of 
Che probability of vaporization in the flow path. 

The pressure-balanced concept selected was conceived by NASA/I. eRC and recom- 
mended tor the LOX turbopump application, because it obviated vaporization 
problems. Additional details concerning', the analysis and design are presented 
in References [l] and [8]. 

5.2 Analytical Approach 

The parametric studies necessary to optimize the seal design fall into three 
categories : 

1. Steady-state fluid-filtr performance 

2. Minimizing any detrimental effects of thermoelast ic distortions 

3. Ensuring acceptable dynamic response to external excitations 



r 5-1 



In ^nneral, tluid-tilm ijeomet rv 4nd pt'rtorm/ince are first optimized. Thermno- 
lastic disiorfions .ire checked to lee that they do not smniticantly attect 
the tluid-tilm pertormance. Variatiuna in materials and exterior geometry can 
be applied to reduce distortionH. Kinally, dynamic response is checked and 
mass and inertia properties ot the seal ring varied to assure acceptable 
response. 

The fundamental reference for the spiral-groove analysis was the work done by 
K.A. Muijdermann [9]. Laminar thecjry was subsequently expanded to include the 
effects of turbulence and inertia at sudden contractions in a manner similar 
to that described in References [b] and [lO]. Klastic distortions were deter- 
mined using a theory modeled upon that described in Reference [l]. 

The computer code was written for spi ral -groove geometry optimization and 
performance evaluation. The ■in.ilytical procedure is summ/irizt'd in Appendix A. 
The final computer code couldl 

1. Optimize geometrical variables on the basis ot stiffness, flow, or 
fluid temperature rise. 

2. Determine the operating film thickness as a function of st^aled pres- 
sure and secondary seal diameter. 

3. Produce film thickness, power loss, circulating flow, leakai^*.' tlcw, 
fluid temperature rise in spi ra 1 -grooves and seal dam, axial and 
angular stiffnesses, natural frequencies, dynamic amplitude ratios, 
and thermoelast ic distortions and stresses. 

5.3 Configuration for the 5Q-mm Seal 

The general groove geometry is shown on Figure 2-13. Note that deep and wide 
grooves are necessary to pump the highly turbulent fluid, because of its high 
effective viscosity. Table 2-2 defines the principal nominal dimensions of 
the pre . iure-balanced design. 



5-2 



Tht original MTI design layout drawing of Che ^O-rntn pressurc-^-ilanced seal, is 
shovm on Figure 2-14. In Che Cote rig, ic will be installed in a back-to-back 
configuration as shown on Che boctotn lefc corner of the drawing. This was 
done to eliminate excessive thrust loading on the test rig Ciirust bearing. 
The outboard seal is the test seal, while the inboard seal is the thrust 
balancing seal. Using the layout (Figure 2-lA}, the seal vendor, Stein Seal 
Company of Philadelphia, Pa., produced assembly and manufacturing drawings. 
There are some variations between Che original layouts and Che final design, 
buc Che subsequent seal vendor drawings govern, and provide accurate informa- 
Cion regarding the final produce. 

The nonrocacing member of Che spiral-groove seal is made from carbon graphite 
(P-5N) and concains the interior high pressure grooving and feed ho!->s. 

The seal rings do not have metal shrouds or interface pieces. The intent is to 
maintain the mass of the seal ring members as low as possible for improved 
dynamic response. 

The spiral grooving is not machined into Che carbon because of pocencial wipe- 
ouc from a high speed rub. The cesc seal assembly is indicaced on SCein Seal 
Company drawing. Figure 5-1*. Details of the carbon graphite face seal are 
shown on Figure 5-2. There are three slots on the outer periphery that mace 
with antirotation pins on the housing elements. The design incorporates 
provisions for a thermocouple installation to measure the feed groove temper- 
ature. Figure 5-3 shows details of the mating ring or spiral-groove runner. 
The spiral-groove lands and dam and the seal land regions were coated with 
tungsten carbide containing a chrome-cobalt binder (Linde LW-15). The base 
material for the spiral-groove mating ring is Inconel 718 to insure structural 
integrity when exposed to centrifugal and pressure force fields. An interest- 
ing feature of the design which was prescribed by Stein Seal Company is the 
circumferential secondary seal, it is of split construction and held together 
by a garter spring. There are two s<>aling surfaces; a stationary radial 
surface between the secondary seal and housing and a sliding sealing surface 



♦Figures are presented consecutively, beginning on page j-ll 



5-3 



b«Cw««n ch« inn«r circunif €r«nce ot the sscondary seal and the secondary seal 
OD on the face seal element. The sliding seal is pressure balanced signif- 
icantly to prevent excessive crush and triction when high pressure is applied. 
The secondary seal is made from the same carbon graphite material as the face 
seal (P-SN). A wavy waihtr spring loads the secondary seal against the hous- 
ing. 

Figure b-U is a drawing of the secondary seal and Figure 5-5 is a photograph of 
the secondary sral- On the left side of the picture, the secondary seal is 
installed in its housing; on the right, the three sectors that comprise the 
major seal components are shown separated from one another. 

5.4 Calculated Fluid-F vim Performance of the 50-mm Spiral-Groove Seal 

A series of computer runs were made over a range ot speeds and pressure. Prin- 
cipal results are tabulated on Table 5-1*. The indicator, C, stands for the 
spiral-groove region and S for the seiiLin)^ ridge. Cooled means the inlet to 
the groove is at a constant prespecified temperature indicating sufficient 
bypass flow through the seal compartment to maintain the constant inlet 
temperature. On the other hand, uncooled means the inlet groove temperature 
is an equilibrium temperature based upon leakage flow and viscous power gener- 
ation. 

Performance information was computed using a constant viscosity of 1.172 x 
10"^ Pa-3 which corresponds to LOX at -WS^C (-279.4° i") and a fluid density of 
1080 kg/m^ (0.039 lb/in. b. 

Film thickness data is shown on Figure 5-6. As expected, the film thickness 
reduces as the pressure goes up and as the speed goes down. The numbers indi- 
cate that rapid liftoff to approximately 3000 rad/s (28,647 r/min) would be 
desirable to avoid excessive rubbing of this seal. The operating film thick- 
ness at 5864 rad/s (56,000 r/min) which corresponds to a rubbing speed at the 
seal interface of 183 m/s (600 ^t/s) is 0.024 mm (0.0009 in.). It is interest- 



♦Tables are presented consecutively, beginning on page 5-41 



5-^ 



ing to note that h ot r«i (jht-throiinh intlow design which h«d been initiallv 
examined, has superior low-ipeed character i st ics. The reason tor this is that 
a large pressure gradient exists across the grooves ot the int low design , .ind 
there are pressure-induced hydrostatic forces between the spiral -grooves and 
the inside seal dam region that provide a substantial litting tor<.« at low 
speed and high pressure. The straight intlow design, however, would fail 
because of vaporization in the interface and was thus eliminated. 

Axial stiffness is shown on Figure 5-^. The stiffness increases with pressure 
and attains an optimum value at 3000 to 4000 rad/s (28,6^8 to 38,197 r/min), 
depending upon the pressure level. As speed increases, the stiffness tails 
off, probably because of the high film thickness at high speeds. Low-speed 
stiffness is very poor, and again reflects the need to become fluid-borne 
quickly and operate at speeds above 3000 rad/s (28,648 r/min). At an operat- 
ing speed of 5864 rad/s (56,000 r/min), maximum pressure of 5.1/ MPa (/50 
psig), the axial stiffness is 81.1 x 10^ N/m (718 x 10^ Lb/in.). 

Circulating groove flow is shown on Figure 5-8. This is not the leakage flow, 
but is the quantity of fluid that is circul.iti-d through the pumping grooves. 
The circulating flow varies inversely with pressure because filrr thickness 
increases as the pressure goes down. At a maximum speed and pressure of 5864 
rad/s (56,000 r/min) and 5.17 MPa (750 psig) respectively, the circulatin)^ 
flow is 2 X 10" m /s (12.2 in. /s). Note the poor circulatioo at low-spee'l 
conditions, which reflects the poor pumping capability, low tilm thickness, 
and generally poor operation at the lower shatt speeds. L.eal'.i^'s or 1 1 ow 
through the sealing dam, is shown on Figure 5-9. The leakage flow is a func- 
tion of pressure differential and tilm thickness. Film thickness effects ar^ 
usually predominant. Note that the leakage curves cross one another. Not 
only is pressure differential and film thickness affecting flow, but there are 
strong effects of turbulence and inertia, especially at the higher pressure 
differential. These effect] cause the crossover of the flow curves. Leakage 
at operating condition of 5.17 MPa (750 psig) and 5,864 rad/s (56,000 r/min) 
is 2.98 x 10"^ m^/s (18.2 in.^/s). 

Power loss curves are shown on Figure 5-10. As expected, pover loss increases 
with speed and pressure. For the pressure-balanced seal, power loss is quite 



5-5 



• ubst«ntt«l because ut the relatively ldrji(t« iniertace area. At the operattm^ 
speed ot 5864 rad/s 06.000 r/min) and 5.17 MPa (750 pHi|() prcisure ditferen- 
tial, the power lots la 9.^7 kW (12.7 ho). 



Temperature rises of the fluid being circulated through the spiral grooves and 
leakinn through the sealing dam are indicated on Figures 5-11 and 5-12, 
respectively. These temperature rises assume that there is external cooling 
(low entering the seal cavity, thus maintaining a constant inl«>t temperature. 
It also presumes that all heat transfer occurs between fluids and not to the 
outside ambient. Exorbitant temperature rises occur at the lower spefd 
(<},00Q rad/s) (28,6^8 r/min) conditions because ot the very low flow to carry 
away the heat generated. At operating conditions of 5.17 MPa (750 psig) and 
5864 rad/s (56,00 r/min) the groove fluid temperature rise is approximately 
JI.A'C (38. 5°?), and 2.83''C (5.1°^) through the seal land. Both ot these 
temperature rises are acceptable. For a noncoole'l inlc cf)nd : t i on , the equi- 
librium temperatures become much higher. These have been calculated. The 
results are a 35.6''C (64.1''F) rise through the spi ral -grooves and a 17. 2''C 
(31*F) rise through the seal land. In addition, the equilibrium grou e 
temperature increases tr'im -IIS^C to -104°C (-ISO^F to -155°K). j) 



5.5 Dynamic Analyais of the 50-mm Spiral-Groove S eal 

Initially, three types ot dynamic .inalyses w-fe considered. The analyses 
started from a simplified approach and th-M graduatpd toward roal-time 
response an.ilyses including Coulomb friction in the secondary seal . In <• , \\ 
analysis, the fluid film was approximated by fluid-tilm stiffnesses obtained 
from the spiral-groove flaid-film analysis. Cross-coupling between axial ind 
angular stiffness was neglected and only principal angular stiffnesses 
applied. Cross-coupling only occurs in the seal land regions and not in the 
spi ral -groove region. Since the spi ral -groove interface is dominant, cross- 
coupling can be safely neglected. In the an .yses considered, the angular 
stiffness was computed on the basis of the computed axial stiffness at the 
operating condition (i.e., assuming an infinite n'-.nber of axial springs). The 
relationship between axial and angular stiffness is as follows: 



'i 2 



LJ 
LI 
l! 
LJ 
(I 
LI 
L 



D 

LJ 
n 



Li 

u 

LI 


D 

5-6 pi 



whcrti 

K9 ■ «ngul«r sctttn«»t 
Kg ■ axitl «c 1 1 1 n«sa 
R ■ mc«n teal radfit 

Fluid film damping w«i« negleclttd. 

Comp«raciv(* restilti of the simplified and c mpreheniive dynamic itudies indi- 
cated that th« differences were not signiticant. Therefore, the iimplified 
methods were applied to tar.iliiate paramt-tric eval aat ions . 

The analytical fflodel for the simpliticd approach considered two circular flat 
plates separated by linear springs. Kluid-film dampini^ and secondary seal 
rin({ friction were neglected. Axi.il stittness ot ih»' springs was obtained 
from the tluid-tilm analysis (see Figure b-7) and angul/ir stiffness obtained 
from Equation 5.1. One plate was vibrated m the axial and angular mndix*. The 
results could be put into a general format for both axial and angular 
vibrat i ohm a-i foil nws : 

AW -2 

Ah . '^ (5.2) 

where: 

Ah = change in clearance 
a "= vibration .amplitude 
U) » ratio ot operating t ruqupncy/nat ur.il fretjuency. 

The quantity Ah/a, defined as the amplitude ratio, can be interpreted as the 
clearance closure at the outer radius divided by one-halt of the TIR runout of 
thp mating ring for angular vibrations. For axial vibrations it is the 
cyclical clearance closure divided by one-halt of the total supplied ampli- 
tude. The angular amplitude ratio as a function of speed and pressure is 
indicated on Figure 5-13. The misalignment angle was 0.75 m rad , '/hich gives 
a peak-to-peak amplitude at Che OD of the seal portion uf Lne mating ring it 
0.0381 mm (0.0015 in.). 



5-7 



Except tor luw-pr«««ur* , hi)(h-4p«ad condi t i onn , the result inn Amplitud* 
r«tiut indicAt* v«ry |ood dytuinic rviponic with l«sii ihan 201 o( ih« available 
cl«aranc« consumed by r«.apons« lag ot th« seal rinn. The hi^h iiiltnens ac 
relatively large clearance* (due to turbulence) contribute to the good overall 
pertormance. The reason tor the poor portormancc at the low pressure, 
high-speed operation is due to very high operating clearances with consequent 
poor stiffness. The seal should not be run at low-speed, high-pressure condi- 
tions, because ot the poor start-up characteristics previously mentioned. 

i.6 Elastic and Thermal Distortions ot the SQ-wm Spi ral -Groove Seal 

Combined elastic and thermal distoriions ot the pressure-balanced design are 
shown on Figure )-l<>. In computing the thermal ettects, it was assumed that a 
percentage (approximately ^Ot) of the fluid temperature riie previously 
computed by heat balance, penetrated the si-al nti^ thickneis as indicated on 
the figure. The thermal penetrations were taken to be rfciangular to a depth 
ot apprunimaiel y 3.i mm (Q.l] in.). This temperature distribution would 
produce greater distortions than actually experienced. A 45-N (10.1-lb) 
sprinn force was .i-i-iumed to be acting on the s.-al rinn in the position shown. 
Also, the secondary seal is approximately 15.75 mm (O.bl in.) t r>m the seal 
face. The distorted position ot the seal rin^ indicates thac it has moved 
radially approximately 0.02b mm (U.OOl in.) and the face has tilted in a 
converging direction approximately 160 Urad. This tilt translates into a 
variation in film thickiiess across the seal face ot only J. 2 x 10" mm (0.l2b x 
10 in.) which is ins igiii t i < ,ini . Since the length beyond the secondary se.il 
is not exposed to the high pressure, it will not move in as far radi' ly as the 
length on the inboard side ot the secondary seal that is exposed to the hi^h 
pressure. The net effect is the bend in the seal ring indicated by the dashed 
line in Figure 5-lA. At its very end, this distortion is apprux imatel y 
0.015 mm (0.000b in.). Thp slop** of this bend must be taken into account when 
designing the secondary seal. The net result indicates that the distortions 
are not excessive and will not havo a serious effect on performance of the 
seal . 



a 
s 

B 
D 
D 



e 



e 

D 
D 
[] 

U 
L 
C 



5-8 



i,l D«tiin of th« 20-tTwi S p i r j I 



-I. r 



Kesulti <>t computerized analyais over <i Hpeed-preHsure rAn^e <ire indtciited on 
Table i-2 4nd on Kinures 5-15 throunh 5-26. Hesulr-i ar*- iiftiilar to t \\¥ 5lJ-rTwn 
'^esiitn in that only prrtinent ronvnrni h .iri> summer i zed belowt 

• Film thickne-ii (es shown on Figure 5-15) varies inversely with pressure 
at expected. At maximiim pr«ssur» the jpeed should be abuvf )bb5 rad/s 
(35,000 r.'min) ' o operate iately. As with the 50-rroTi dcs ign , Kjw-spoed 
performance is very p<Jor. 

• Axial stiffnesr is shown on Kinuro 5-16. I.ow-speed stiffness is poor. 
At maximum speed and pressure, th«? stittneis is 48.5 x 10 N/m (277,000 
Ib/in.) 

• Leakage flow is indicated on Figure 5-17. At maximum speed and pres- 
sure, the leakagt! is 1.9 x 10~ m/s (11.6 in. /s). 

• Circulating groove flow is shown on Figure 5-18. At maximum condi- 
tions, the flow iy 1.38 x 10" m/s (8.4 in. /s). The maximum circulat- 
ing t 1 ow occurs at low-pressure, high-spi^t'd ccjnd 1 1 i ons , because ot the 
hi^h film thirUnPS'i .it this condi. ion. The maximum flow .s 1.5') x lO" 
m/s (9.5 in. '/s). 

• Powpr loss is indicati'ii un Figur»> 5-19. Maximum pow^r is dissipated at 
maximum preiiuure .ind speed and is equ«' ^o 7.084 <^ (9.5 hp). 

• The temperat un- risi' in the seal land is shown on Ki^ure 5-20. The 
20-mm seal was only analyzed at the cooled inlet condition since that 
is the anticipated operation. The maximum tei,-'|>"rat ur« ris'; through the 
seal i «. only 2.28''C (^.l^F). At the ..ower speeds, over-temperature 
would occur at sustained operation. 

• The temperature rise ot the c i rcul at i n>>. Moid thr<nigh the grooving is 
ahown on Kigure 5-21. At maximum speed and pressure, the temperature 
of the circulating tluid rises approximately 25**C (AS'F). 



5-9 



• The n.itural frequency of rhe leal rinn in the axial mode is shown dp 
I'ltiure ^-22. It is only at the low-pressure conditions that the 
trequency approaches uncomt ort abl y close to the operating speed. 

• Angular natural frequency is shown on Figure 5-23. Again, the indi- 
cations are that low-pressure, high-speed operation should be avoided. 

• Axial ainplitude ratio is shown on Figure 5-24. It clearly indicates 
the danger tor operating below a pressure ot 1./24 MPa (250 psii^) at 
full speed. 

• Angular amplitude ratio is shown on Figure 5-25. Identical cotnments as 
for the axial amplitude ratio apply, 

• Figure 5-26 shows elastic distortions. These are all very tnoderate. 
Elastic distortions produce a variation in film ihickness of 0.0008^ cm 
(33 Uin.) which is negligible compared to the op<!rating film thick- 
ness. 



5-10 



OE POOR QU*«J^ 



FUU TMSfAO UJPTm 

TAUG RfliOvr. 

MIN BOLI POOJlCTION i8C 




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I (i)thebmocouple — ' 

INSTALLED IN fACE 

SEAL 



End view 



i:ii:^<>\yi iiovNC 



Fig. 5-1 50-mm LOX Spiral-Groove Seal Aa 






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5-13 



851648 




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WIN JiNSiLC (nCLO o2%orrsrr) 

NOTTS. 



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50MM SIZE 



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50-mm Spiral-Groove Seal Mating Ring 



5-15 



85:647 



ORIGINAL PA^'£ J5 
OF POOR QUALITY. 



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SO-rrm Spi ral-Croove Secondary Seal 



5-17 



85 1644 



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■4- 



1000 2000 3000 4000 5000 6000 7000 8000 

Speed, rad/s 



Fig. S-b 30-Ma Spiral -<;roove Seal; Fila riUikiiess Ver:>u^ bjiicd and I'ltssurt 



9»t» 



I 



X 

E 
E 






<0 

a 
x 

< 



5000 



7 MPa 



= 3 45 MPa 




7000 8000 



Speed, rad/s 



33511 



Fig. S-7 50-BB S|)ir.il-('.rixtve Seal; Ax i t ! Stittiiess Versus Speed and Pressurt.- 



I 



X 



O 



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> 

o 
o 

6 




4000 
Speed, rad/s 



8000 



iJS^Z 



Fig. 5-8 SO-aw Sp lr;il -Croove Seal; Spiral (;ruovi- tireulalin); Flow Versus 
Speed and Pr<-s&iirt.- 



Eir: a rn = 



I 



X 



O 



0) 

03 
0} 




3000 4000 5000 

Speed, rad/s 



8000 



Fig. S-9 50-aM Splral-iJroove Seal; Leakage Flow VerMts Spe«d aitd PressMrc 



I 




P = 5 17 MPa 
P = 3 45 MPa 
P = 1 72 MPa 
P = 69 MPa 



1000 2000 



3000 4000 5000 
Speed, rad/s 



6000 7000 8000 



Vi^. 5-10 3U-mm Splrj 1 -(.move Sta I ; Power Loss Vt-rsii.-, Sptcd and Pressure 



aj6i4 



C3 C3 C C 



t i 



-5.17 MPa 



I 



o 

o 
0) 

CO 

DC 
a> 

D 
OJ 

CL 

E 



0) 

> 

o 
o 

6 



Cooled inlet 




8000 



Speed, rad/s 



:j6i5 



Fi>;. ')-ll Sd-inni Spir.il (;roi)ve Seal; Sp i r.i 1 Kiroovc Teropor.jture 
Kisf \'>'rsiis Speed and Pressure 



P - 3.45 MPa 



Cooled Inlet 







2000 



3000 4000 5000 

Speed, rad/s 



7000 8000 



13616 



Fl^r. 5-i:- Sn-mm Spiral -Croove Seal; Temperature Klse Versus Speed .md F'ressn 



re 



I mi l—^lt i I 



• t 



aDricziczrccr'ao 







2000 



3000 4000 5000 
Speed, rad/s 



6000 7000 



P = 3.45 MPa 
P = 5.17 MPa 



8000 



3i619 



Fi^. 5-13 50-iiim Sp i ra 1 -Cr<.»>vo Seal; Angiilir Ampl 1 tmlf K.it In Versus Speed and Pressure 



00 



28 



24 



20 



t5 16 



X 

E 
E 

c 
o 

o 
b 



(0 

X 12 
< 



E 
E_ 

c 
o 

(0 

o 

Q. 

X 

< 



8 

4 
-4 

-8 




8 = 0.026 mm 



Secondary Seal 



J- 



^ 08**C 




45 N 
Spring 
Force 



laa^'C 



J L 



1 



1 



J L 



J L 



J L 



1 



4 8 12 16 20 24 28 32 

Radius (mm) Radial Distortion (mm x 10^) 

KiK. i)-14 "jlJ-mni Spiral -tiroove Seal; rhcTim»fl.ir.t ii DibtDit ion 



36 40 



J 
44 



tUIIJ 



fc^^ C! c rr! ca : 



■ « ■ 7 w ■% 



I 



0.06 



005 

I 

CO 004 

CO 

« 

c 

o 

j5 0.03 

E 



0.02 



0.01 




20.000 40.000 



60.000 
Speed (rpm) 



80.000 100.000 



I I i I I I i i_ 



_l L. 



2.000 4.000 6.000 

Speed (rad/s) 



8.000 10.000 



^"i^;• J-l') Ml-imii S|iii.il Cro.ivc St.il, I" i 1 m llii.ktu-ss v»•rs^l^. Spt-i-d 



BS1618 



o 





/.u 




40 








3.6 




6.0 




32 


1^ 




tf> 




o 


50 


o 


28 


X 




X 




E 




c 


?4 


z 


40 


— ~^ 




■^—^ 








8 

0) 

c 


3.0 


CO 


2.0 


«: 
S 




Si 


1.6 


.5 

X 

< 


20 


5 

X 

< 


12 
08 




1.0 




04 







_ 






u 




P = 5 17 MPa 




20,000 40,000 60,000 

Speed (rpm) 



J L 



80,000 100,000 



J i_ 



2,000 



4,000 6.000 

Speed (rad/s) 



8.000 10.000 



Klj;. 'j-lh JO-mni S|>ii il i.ionv. Si-.il; ,\.\i.il Stitliu-^is vi-rsti^ SiKr\-d 



U16I7 



cr: trz 3 



"I w 



I 



X 

CO 



O 



0) 
O) 
CO 

CD 
0) 



20 
18 
16 

14 
12 

10 

8 

6 

4 

2 
0- 



12r 




20.000 



40.000 60,000 
Speed (rpm) 



80.000 100.000 



I I L 



' ' 



2.000 4.000 6.000 

Speed (rad/s) 



8.000 10.000 



Ki};. i-l / .'o-mm Sjiir.il Cnuivi- SimI; l.i-.ik.i^t' Klow vi-r. is Spt't-d 



asi6n 



I 



16 

14 

o 12 

X 

- 10 
8 



o 



> 
o 
o 



6 
4 
2 
0- 



co 



o 



0) 

> 
o 
o 

O 




L. 





20.000 



40,000 



60.000 
Speed (rpm) 



MPa 



80.000 



100.000 



2.000 4.000 6.000 

Speed (rad/s) 



8.000 



10.000 



I- i){. >~\ti Jii-Mu Spii'.il CriHivt.- S«.-al , (^riMivt- Kiow vt-rsus Spt-oti 



•S1607 



tx: i— □ 



CaC3ir3C3C2aCDCD 



r*? n 








7 




6 


$ 

JC 


5 


(A 
CO 

O 

_l 


4 


o 

Q. 


3 




2 




1 








Q. 

(A 

o 



0) 

o 








20,000 



40000 60.000 

Speed (rpm) 



80.000 



100.000 



2.000 4.000 6.000 

Speed (rad/s) 



8,000 



10.000 



^ig• >-!** lO-mm Si>ii.il CnHtvL' Sf.il; I'ow. i l.o.ss vt•r^>»l^. h|»«.-eu 



tsteis 



I 



O 



I 
I 

I 



13 

12 

11 

10 

9 

8 

7 

6 

5 

4 

3 

2 

1 





P = 5.17 MPa 



0) 
(0 

ir 

0) 

D 

a 
a> 

C 



CD 
CO 




20,000 40.000 60.000 

Speed (rpm) 



80.000 



100.000 



2.000 4.000 6,000 

Speed (rad/s) 



8.000 



10.000 



^■i:i• *--'** -'U-B*i S|»ii.il <;iiKiv%- Sf.il; S*-.il l'<M|H r.iiiirt- Kit**- vi-rMis S|K-i-d 



•61616 



CZ3 cn □ cn 



C2 □ □ C3 CD E3 



300 



o 

o 

2 200 
a: 

2 150 

a> 

a 

E 
a> 

t- 

(D 100 

> 
o 
o 



o 



50 



500 




20,000 



40.000 



60.000 



80.000 



100.000 



Speed (rpm) 



2000 4.000 6000 
Speed (rad/s) 



8.000 



10.000 



^''g• ■>--• 20-mm S|>ir.ii Cumivc- S*-.il; S»-.il li aiH-i.tlurf KIm- vtr.SMi> S{h-«>«J 



K160* 



I 



50.000 



■ 40.000 



^ 

s 



3 

5 30.000 



m 
\- 

3 

Z 20,000 

a 
"x 

< 



10.000 



X 

E 



500 



400 



i 300 

3 

o- 

0) 



5 200 



< 



100 







1 



P = 517 MPa 




J L 



J I I L 



-i 



20.000 40,000 60.000 
Speed (rpm) 



80.000 100.000 



J L 



2.000 4,000 6.000 

Speed (rad/s) 



8.000 10.000 



Fij;. i-11 ^(»-aM S|> i i .i I i.rtMtvt- S*-.«l; Axi.il N.iiiir.il Kr«.-<]iK-it> v vt-rNiis S|Kt.-«i 



•S1612 



ac3C3C3C3l=IQ5aC3C3CJCD 



I 



•30.000 



■g 50,000 



u 

§ 40.000 



a 



3 

at 
c 
< 



30.000 



20.000 



10.000 







600r 



° 500 

X 

E 
a 



400 - 



u 

c 



3 

a 

5 300 



(0 

3 
CO 

Z 



200 



3 

S 100 



P = 5l7MPa 




_L 



20.000 



40.000 



60.000 



80.000 



Speed (rpm) 



100.000 







-I I- 



2.000 4.000 6.000 

Speed (rad/s) 



8.000 10.000 



Kig. '»-_' I _'»>-■» S|iir.i( Croovt- S«-.il; AiiKul.ir '<.itiir.i! Fr^qut-utv w».-rsMs S|*«-«-«l 



•SMI) 



r 



cc 

0) 
XJ 

E 
< 

.5 
"x 
< 




20.000 



40,000 60,000 
Speed (rpm) 



80.000 100.000 



■ ' 



' ' L. 



2,000 4.000 6.000 

Sp>eed (rad/s) 



8.000 10.000 



Ki>;. J-J4 Jil-imii .S(>ii.il i.i.mvi Si.il; Axi.il Amp I 1 1 u'i. K.ii ii> viTsiis S|»i-«-il 



•51614 



a 



□ □cucacccaninocan 



r 




20.000 40.000 60.000 80.000 100.000 

Speed (rpm) 







2.000 4.000 6.000 

Speed (rad/s) 



8.000 



10.000 



Kig. S-J*) J()-nuii Spiril Croovi St .i I ; Aii-ul.ii AmiilimJ. K.iti.i vi-r.su;. S(uiJ 



S61609 



I 

O 



20 



E 15 



I 

I 10 
a. 







i 

« 
u 
ea 

a 

a 
t 

« 



c 
o 

s 

< 



u.» 


' 


0094 


fnm 














A = 0086 mm 


0.8 


I •" 


/ 




0.7 


\ 


1 / 
I • 
I • 

'' 1 ^ 
/ 1 ^ 

1-'"' / 


/ 
/ 
/ 

/ 
/ 




secondary Seal 
8N 


0.6 


494 


O.b 
04 






Spring Force 






I' V 








0.3 




^ 


■ 


' 




0.2 










1 tt ^ 86M-rad 




1 




0.1 

n 




1 -n 


1 il 


1 

.1 11 . . . 


1 1 



0.2 4 6 0.8 10 1.2 

Radial Position — in. (Relative Displacement — mils) 



_L 



X 



X 







10 IS 20 

Radial Position (mm) 



25 



30 



^i^- >--'»» S|ti 1 .1 1 -CrMovf Sill - t I isi i. Di >,( t>rt i»iis 



B5I610 



□ □ C3 f!!3 



r^ □ ra n c*! 



TABLE 5-1 



50-m PRESSURE-BALANCED SEAL 



(rad/s) 




Pressure 
(MPa) 


FiSM 

Thlcknass 

inm) 


(a) (s) 


PoMSr 

Loss 

(hM) 


Temparatur* . 

Cooled 

(-C) 

(g> (s) 


Taaparatura. 
Uncoolad 

{-O 
(g) (s) 


AaisI 

Aaplltud* 

Ratio 


Angular 

Ai^tlituda 

Ratio 


Aaial 
Stfffftass 
N/« ■ 10~* 


Faca 
Load 

(N) 


7330 


S.I7 


29.7 


27.5 


39.0 


16.36 


26.7 


3.B 


4S.5 


22.6 


0.042 


0.031 


7.38 


lo.eoa 


5864 




23.7 


20.0 


29.8 


9.46 


21 .4 


2.8 


35.7 


17.1 


0.024 


0.018 


8.07 




5236 




20.98 


16.6 


25.6 


7.23 


19.8 


2.5 


32.5 


tS.2 


0.019 


0.014 


8.37 




4169 




16. 1 


10.5 


17.9 


4.25 


18.4 


2.1 


29.2 


12. B 


0.011 


O.OOB 


B.70 




3142 




10.44 


4.6 


9.34 


2.24 


22.3 


2.0 


33.2 


12. B 


0.007 


0.005 


B.26 




2199 




2.79 


0. 16 


0.510 


1 . 12 


300.6 


21.3 


401 .0 


122.0 


0.013 


0.009 


2.18 




7330 


3.15 


36.3 


29.3 


39.9 


15.3 


23.3 


3.61 


40.6 


20.6 


0.072 


0.0S3 


4.<6 


7.086 


586'-. 




29.7 


22.3 


31.5 


B 8 


17.7 


2.57 


30.0 


15.0 


0.040 


0.030 


4.96 




5236 




26.4 


19.0 


27.5 


6.6 


15.7 


2.18 


26.7 


!5.8 


0.0303 


0.023 


5.21 




3142 




14. e 


7.21 


12.7 


1.94 


12.4 


1 .32 


19.4 


8.3 


0.009 


0.007 


S.8I 




1760 




2.54 


0.098 


0.272 


0.67 


308.9 


24.1 


417.0 


132.0 


0.016 


0.012 


1.12 




7330 


1.72 


49.5 


30.97 


39.2 


14.1 


20 . 26 


3.4 


36.1 


19.4 


0.188 


0.135 


1.69 


3.562 


5864 




41 . 1 


24.42 


31 .8 


7.9 


14.4 


2.4 


25.6 


13.3 


0.099 


0.072 


2.14 




5236 




37.3 


21 .47 


28.5 


5.9 


12.2 


1.9 


21.7 


M.I 


0.072 


0.053 


2.28 




4169 




.3.06 


10.65 


15.8 


1.64 


7. 1 


0.94 


1 1.7 


5.6 


0.020 


0.015 


2.81 




2094 




14.22 


4.42 


7.93 


0.67 


6.6 


0.72 


10.6 


4.4 


O.OOB 


0.006 


2.91 




1309 




3.73 


0. 179 


0.456 


0.30 


65.6 


5.4 


91. 1 


31 .1 


0.011 


0.008 


0.860 




7330 


0.69 


70.36 


31 .22 


34.37 


13.04 


Id. 6 


3.7 


35.0 


20.6 


0.844 


0.524 


0.65 


1.450 


5864 




59. 18 


25.25 


28.84 


7.23 


12. 7 


2.4 


23.9 


13.3 


0.347 


0.240 


0.75 




5236 




54. 10 


22.58 


26.06 


5.36 


10.5 


2.0 


19.4 


11. 1 


0.23B 


0.169 


0.79 




3142 




35.31 


12.95 


16.06 


1.42 


4.9 


0.83 


8.9 


5.0 


0.055 


0.041 


1.05 




2094 




24. 13 


7.37 


9.87 


0.52 


3.3 


0.47 


5.6 


2.8 


0.020 


0.015 


1.22 




1047 




8.69 


0.97 


1.74 


0. 15 


5.4 


0.57 


8.3 


3.3 


0.007 


O.OOS 


0.912 




890 




2.79 


0.46 


0.0<>2 


0.07 


82.8 


0.011 


125.0 


50.0 


0.G20 


015 


0.228 





TABLE 3-2 



20-M( PtESSURE-BALANCED SEAL 






r" ~ 



N 
(r«d/s) 


Prassura 
(MP*) 


Fllni 
Tnichnask 

(IM) 


PoMcr 
Loss 


(g) (•) 


T«aip«r«tura. 

Coolao 

(-C) 

(g) (s) 


AiiUI 

Ratfo 


Angular 

Aa«iMtud« 

Matt a 


Aaial 
SttfffnasK 
H/ai ■ I0~^ 


Faca 

Laad 

(M) 


10.«72 


5. 17 


25.91 


7.08 


138.0 


1 90. 


;>4.8 


2.3 


0.073 


O.OS1 


4.65 


6.363 


7.854 




19.56 


3.51 


88.5 


134.0 


18.9 


1.6 


0.03S 


0.025 


5.42 




6.545 




16.0 


2.31 


63.9 


103.0 


17.2 


1.3 


0.023 


0.01* 


5.67 




5.236 




1 1 .94 


1 .34 


37.7 


68.8 


17.5 


1 . 10 


O.OIS 


0.010 


5.66 




3.297 




7.11 


0.745 


12.78 


29.7 


27.2 


1 .3 


0.010 


0.007 


4.75 




3.126 




2.54 


0.522 


0.98 


2.62 


256.0 


12. 1 


0.021 


0.015 


1.40 




10.473 


3.45 


31.50 


6.63 


146.0 


192.0 


21.8 


2 2 


0. 123 


O.OSS 


2.96 


4.270 


7.854 




24. 13 


3.28 


102.0 


1 «1 .0 


15.4 


1.39 


0.057 


0.040 


3.42 




6.545 




20.07 


2.09 


77.0 


113.0 


13. 1 


1.09 


0.037 


0.025 


3.65 




5.236 




15.75 


1 . 19 


50.8 


81.9 


11.7 


0.87 


0.022 


0.016 


3.62 




3.927 




10.92 


0.b7l 


24.6 


45.9 


12.7 


0.77 


0.013 


0.009 


3.72 




?.5ei 




2.54 


u.2i)8 


0.66 


1.80 


206.0 


10.0 


0.026 


0.016 


7.69 




10.472 


1.72 


41 .91 


6. 1 1 


154.0 


183.0 


19. 1 


2. 13 


0.327 


0.213 


1.33 


2.157 


7.854 




33.02 


2.91 


112.9 


134.0 


12.5 


1.32 


0.135 


0.093 


1.56 




6.545 




28. 19 


1 .86 


90.0 


115.0 


9.9 


0.99 


0.0*2 


0.057 


1.70 




5.236 




22.86 


1 .04 


66.2 


89.5 


7.8 


0.72 


0.046 


0.033 


1.66 




3.927 




17.02 


U.54 


40.6 


60.0 


6.4 


0.53 


0.024 


0.017 


2.00 




1 .932 




2.54 


0. 15 


0.459 


I. 10 


IJ9.0 


5.94 


0.029 


0.020 


0.4O 




10.472 


0.69 


56.90 


b. 74 


156.0 


152.0 


17.5 


2.39 


1.7 


0.82 


0.522 


690 


7.854 




45.47 


2.68 


1 18.0 


1 W.6 


10.9 


1 .44 


0.43 


0.27 


0.€i2 




6.545 




39.37 


1 .64 


97.0 


100.0 


8. 17 


1.06 


0.23 


0.16 


0.677 




5.236 




32.77 


0.97 


75.4 


80.3 


5.94 


0.72 


0.12 


o.ce4 


0.760 




3.927 




25. 15 


0.4b 


52.0 


59.0 


4.22 


0.49 


0.057 


0.040 


0.661 




1.445 




2.79 


0.52 


0.33 


0.49 


70.0 


5.8 


0.033 


0.023 


0.196 





r. c:! rr! rr: rr m m 



En n rr: rr: 'T? a n 



6.0 TEST RIG 

6.1 General Configuration 

A cross section of the test rig is shown on Figure 6-1*. The right-hand 
portion of the rig is the drive end where the nitrogen turbine is located. The 
central portion is the bearing region where the journal and thrust bearings 
are located. The left end of the rig is the test seal section where the 50-mm 
helium buffer seal is shown installed. 

Locating the thrust bearing in the center of the rotor avoids excessive over- 
hang at either end and provides for a more uniform distribution of mass along 
the rotor. This arrangement alleviates rotordynamic difficulties due to large 
overhung masses. The helium buffers are installed in a back-to-back coni'ig- 
uration and mate against a common runner. 

The 30-mm (1.1811 in.) shaft journal diameter provides sufficient stiffness 
to be below the bending critical speed, and prevents excessive bearing and 
windage power losses for operation at 7330 rad/s (70,000 r/min). At the 
turbine end of the shaft, a heat dam is located between the turbine wheel and 
shaft. This dam prevents high temperature at the turbine wheel from conduct- 
ing heat into the cold shaft regions. The outside periphery of the heat dam is 
machined with a labyrinth that provides one half of a buffer seal that 
prevents turbine gas from entering the bearing region. At the seal end of the 
shaft, the helium seal runner is secured to the shaft by a spring sleeve that 
is pressed onto the shaft. This compensates for bore growth of the runner. 

Photographs of the seven major housing components are shown in Figures 6-2 and 
6-3. At the turbine end is the nozzle box which contains the turbine inlet 
manifold, nozzle blades and exhaust section. The turbine blade tip clearance 
of 0.25 to 0.30 mm (0.010 to 0.012 in.) is established by the width of the 
spacer piece between the nozzle box and labyrinth seal housing (see Figure 
6-1). The labyrinth seal housing is a separate member that contains inlet and 



^Figures are presented consecutively, beginning on page 6-lb, 



6-1 



exit passagea tor the buUfir gas. The next major housing element: is t hf 
turbine end bearing housing. This contains the turbine-side thrust and jour- 
nal bearings with associated fluid inlets, drains and pressure tap 
connections. The shaft support journal and thrust bearings are externally 
pressurized (hydrostatic) bearings to provide stiffness and avoid whirl 
instabilities. The bearing housing also contains the right-hand side of the 
windage shroud surrounding the thrust collar. Separating the two bearing 
housings is a shim plate whose thickness determines the total clearance in the 
thrust bearing. The seal end bearing housing contains the left-hand thrust 
and journal bearings. The n?xt major element is the seal housing in which the 
test seals are installed. Finally, the end cap is the last member and is used 
to contain the fluid that leaks past the tast seal. (Photographs of rig 
components and an assembled shaft are included in Section 2.0, Figure'j 2-20 
and 2-21.) 

The assembly of the rig proceeds from inboard out. The bearing Housings and 
shi-n plate are installed over the shaft and connected by cap screws torqued to 
their specified values. Next, the assembled buffer seal housing is installed 
and connected to the turbine side bearing housing. The heat dam is then 
shrunk onto the shaft. The turbine nozzle spacer is next inserted. The 
turbine wh 'el is chiUed in LN2 and inserted over the shaft into the bore of 
the heat d.im. A washer and nut finish the shaft assembly. The final step at 
thf. turbine end is installation of the nozzle box. At the opposite end the 
helium seal housing is installed. The seal runner spring sleeve assembly is 
then attached to the shaft and the seal runner shrunk on and secured by the end 
nut. The test seals are then put in place followed by the end cap. 

6.2 Fljjid Sy stems 

The variour, flow systems and flow paths in the rig are indicated schp'.atically •»■ 
on Figure 6-4. The nitrogen gas for driving the turbine enters the nozzle box 
on the right-hand side, proceeds through the nozzle blading, through the radi- 
al-inflow turbine that drives the rig, and out through a central exhaust in 
the turbine nozzle box. Leakage at the back side of the turbine wheel passes 
through a labyrinth seal ,' mixes with the buffer seal gas, and exits from a 
common drain. , he buffer seal consists of a central supply annulus from which 



6-2 



tlow in directed axially through labyrinth'} toward^i the turbine whetM .mU in 
the opposite direction towards the bearing compartment. 

Initially, nitrogen gas was used ifor the buffer tluid in the labyrinth seals. 
The check-out test showed, however, that the temperature oi the surroundin*? 
parts was low enough to permit some of the nitrogen to liquefy. This ran back 
into the turbine posing a problem. Helium gas with its much lower liquefica- 
tion temperature was substituted for Che nitrogen and solved the problem. 

The journal and thrust bearings are externally pressurized by liquid nitrogen. 
There are four recesses in each bearing, each fed through an ext«rnal passage 
via an orifice restrictor located at the outer periphery of the rig, The 
drain from the turbine end of the turbine side journal bearing mixes witn the 
labyrinth seal drain. The inboard drain from Che turbine side journal bearing 
combines with the thrust bearing drain located around the outer periphery of 
the central section of the thrust bearing. The inboard drain from the seal 
side journal bearing flows through a separate drain outlet and combines with 
the leakage from the thrust bearing seal drain. 

Two different arrangements were used to drain the outboard side of the seal 
end journal bearing and the inboard helium buffer seal. The original design 
(shown in Figures 6-1 and 6-4) uses a common drain cavity. The helium-nicro- 
gen mixture exited through radial drain holes directly under the journal bear- 
ing. This arrangement, used during the te'icing of the firsc three seal set.i, 
allowed the inboard end of the seal runner to be bached in LN2 resulting in 
large windage losses. The heat generated caused the adjacent journal bearing 
case to warm up and partial vaporization to occur in Che bearing. The vapori- 
zation of the bearing fluid caused deterioration of its stiffnes.s and damping 
properties and resulted in serious dynamic problems at speeds of approximately 
50,000 r/min or greater. 

The second arrangement designed to prevent the vaporization problem uses a 
labyrinth seal to keep the LN2 bearing flow away from the end of the seal 
runner. The configuration was shown in Figure 2-22. The helium side of the 
labyrinth is kept at a higher pressure than the LN2 sidt', thus preventing any 
LN2 from flowing in and contacting the end of the runner. This eliminates the 



6-3 



windftgu UiHH altogether. The h»'l jum is supplicti by the leakage trom t hi* 
inboard test seal and flow irom an additional lupply part machined into thi* 
seal housing. A constant buffer flow is maintained across the labyrinth by 
keeping the externally supplied helium at a pressure slight ly higher than th« 
bearing drain. 

The various fluid supply and drain systems require a large number of passages. 
These are shovm on Figures 6-5 through 6-7. Section numbers are referred to 
in Figure 6-1. Figure 6-5 shows the turbine side bearing housing inlet's and 
outlets. Section D-7 is in the journal bf'aring region and shows the inlet; 
ports to each of the recesses (Connection A), The restrictor element'} are 
located at: the outer periphery of each entrance connection so that they are 
readily accessible for change, if desired. The two bottom recesses also have 
pressure tap Connections B. Inboard of the bearing, at its drainage manifold, 
are capacitance probe Connection.s G, for monitoring shaft motion. Although 
only one probe is shown, there are actually two, located 90 degrees apart. 
Section D-4 is in the right-hand bearing housing in the vicinity of the thrus'. 
bearing. Each of the four thrust bearing recesses are separately fed through 
Connection A, The two top recesse?; have pccn'Uifd t.api B, for me.Tiurink? t;hi» 
thrust bearing recess pressures. There are also four separate Connection;; C, 
for inlets to the vibration pistons [3]. These were not used during the heli- 
um seal test. The several drain holes shown communicate i.he inboard journal 
bearing drain with the thrust bearing drain. 

Figure 6-6 shows supply and drain connections for the seal end bearing hous- 
ing. Note that the left and right-hand bearing housings are not antisymmetric 
because of the different draina^jC systems at the outboard end of the journal 
bearings. Section A-7 (thrust bearing housing) shows the orthogonal capaci- 
tance probes axially positioned on the inboard side of the journal bearing. 
Figure 6-7 shows the connectionr for the turbine end labyrinth seal and shim 
plate housings. The labyrinth seal housing. Section D-IO, contains one inlet 
Connection A. It is at the bottom and feeds the central region of the double 
labyrinth seal (see Figure 6-1). At the top of the housing. Connection B, is 
the combination drain for the LN2 from the outboard end of the turbine-side 
journal bearing and the buffer gas leaking past the inboard labyrinth. 
Connections C exhausts combined leakage from the turbine and outboard 



6-4 



labyrinth -j^al . The 'ihim plate, Section A-'), hai tour drain cunnect luns 41 
the bottom which are nutlets tor the thrutt bearing and the inboard leakax^? 
from both journal bearings. 

^. '.1. ,. J® St Seal Arrang emenc 

The helium seal arrangement is shown in Figure 6-1. The inboard seal contact*; 
the mating surface o£ the seal housing. The outboard leal contact's a similar 
mating surface and is retained by the end flange which is internally bolted to 
the seal housing. The seal rings are prevented from rotating by Lhrep equally 
spaced, key-like protrusions machined into th« bore, ot the yeal housing and 
end flange. The seals are pushed apart providing a small, but positive, axial 
sealing force by a wavy washer installed between the rings. 

The helium seal runner attachment to the ."shaft wa-j prnviously shown on Figure 
4-6. A flexure sleeve is pressed in the bore of the seal runner. The assembly 
is then pressed over the shaft end. The purpose of the flexure is to compen- 
sate for centrifugal growth of the runner. As was indicated on Figure 4-6, 
the growth at the ID of the runner ig approximately 0.0104 mm (0.0004 in). A'l 
the runner expands, it allows the flexure to release and maintain contact v^ith 
the runner. 

6.4 Tur b i ne__De s i gn and Pe r f rmanc e 

The turbine used to drive the rig is a radial-inflow turbine detii^ned for a 
maximum of 74 kW (100 hp). The principal losses in the rig are due to windage, 
bearings, and seals viscous losses. Computed total power loss tor the tost 
rig with helium seals installed and the original inboard seal drain arrange- 
ment is approximately 31 kW (41 hp) at 7329 rad/s (70,000 r/min). With the 
labyrinth seal installed between the helium seal and journal bearing, the 
total computed loss drops to about 17 kW (23 hp). Figure 6-8 shows the 
locations of the various losses. 

Figure. 6-9 presents the turbine configuration} Figure 6-10 provides a photo. 
Design and performance data are given below: 



6-5 



Basjr rurbtnt? IJe. .^ 

Type! Unshroiided, cAnt i l#'^eri'd, radial "intlov 

OU, itOT (in,)s 7], 9 (2.8J) 

SpeKd, r«d/3 (r/min)j 7,329 (70,000) 

Tip Spe#d, m/i (tt/s): 284 (865) 

Nozzle Angle, rad (deg): 1.29 (7A) 

Blade Angl«, rad (deg): 1.05 (60) 

Reaction (%)'. 30 

U/Co: 0.5 

Estimated Performanne 

Ga'i! N2 at ai^C (70%', 360 "H 

PQw«r, kW (hp): 37 (50) 

Inlet Pressure, kPa absolute (psia): 1034 (150) 

Exhaust Presjjure, kPa absolute (psia)! 103.4 (15) 

Isen. Enthalpy Drop, Btu/kg (Btu/lb): 27.2 (60) 

Kiiirimcy, :: 70;: 

N2 Flow Rate kg/a (•scim): 0.40 (700) 

OH-design performanue data are provided by Figuro'! 6-11 and 6-12 ■jhowini'; 
output power versus inlet pressi.rp and turbine Ht'tici»>ncy ver-ius •Jp^'t'•i, 
respectively. 

6.4.1 Acceleration 

The specification calls for an acr^ieration rate of 152 m/s (500 fc/s"). 
This is equivalent to 15.5 g. The turoine design speed torque is: 

T = (50)(63, 000)770,000 = 45 in. -lb = 5.08 N/m 

The stalled torque for this type of turbine is two times the design torque, 
10.2 N/m or (90 in. /lb). The 2.15 kg (4.75 lb) rotor has a mass moment of 
inertia of 0.00068 N/m-s^ (6 x lO"*"^ lb/in. -s^). The average acceleration is 
then: 



6-f) 



g • (90 + 45)/l2 X n.006 X J86) « 29 

Thtretort, thert j i ampli* torqutf av4ii4t)l« tu me»t ithts f peei t ic«t »un 4ct*lier'- 
«tif)n r«t«, 

6.4.2 btresi 

At cht design conditions ot 37 kW (iO hp) «t /329 rad/'i (70,000 r/min), th«a 
maximum stresi in 137.9 MP« (20,000 psi) which occurs «t th« bUd« ID t rum 
both IM centritugal forces and torque load. Aluminum alloy 2024-T} has «i 
yield strength ot 413.6 MPa (60 psi), hencs the maximum salu operating spi».'d 
is about 10,470 rad/s (100,000 r/min). 

6.4.3 Desitjn 

The design drawing ot the turbine wheel is shown in Figure 6-13. 

6.5 Bearing Design and Performance 

6.5.1 Journal Bearings 

The journal bfirtring is a tour-p.-id hydro'Jtat i c bearing with a pocket or rect'';-; 
geometry as 'jhown in Figure 6~14. The rece.'i-i angular extent i'i 1.26 rad (72'') 
and approximately 16.4 mn (0.644 in.) wide, ihe bearing width ir> 22.0 mm 
(0.866 in.) and the L/D ratio i'i approximately 0.73. The design radial clear- 
ance is 0.0191 to 0.0254 mm (0.000/5 to 0.00100 in.). The actual mea<5urr.d 
radial clearance was 0.0188 mm (0.000/4 in.). 

Bearing performance is summarised in Table 6-r-'''. Since the program originally 
called for Che use ot I.OX (for spiral groove seal casting) and test speeds of 
10,470 rad/g (100,000 r/min) for testing 20-mm seals, bearing performance for 
those conditions is also given. The data applies to the concentric shaft 
position which, given the low rotor weight, is Che expected mode of operation. 

■■•'Tables are prersented consecutively, beginning on page 6-49. 



6-7 



The bearing flow is very sensitive co clearance and only sLighdy affected by 
speed. The flow is used to sise the orifice to produce a pressure drop between 
Che supply and bearing recess equal to about one half the total drop between 
Che supply and drain. The original orifices installed were 0.9A0 mm (0.037 
in.) in diameter. These were changed during check-out operations to 0.813 mm 
(0.032 in.) in diameter, yielding a recess pressure 0.4S to O.SO times the 
supply pressure. 

Both stiffness and damping are also sensitive to clearance. The bearing land 
areas are silver plated from O.OS to 0.10 mm (0.002 to 0.004 in.) to reduce 
rubbing friction with the shaft. The bearings are energized prior to start-up 
of the shaft and after it has stopped rotating during shutdown. Thus, there 
should not be rubbing contact between the shaft and bearings. 

6.5.2 Thrust Bearings 



I 

I 
1 
I 
I 

L 

t 

t 

L 

L 



The general configuration of the thrust bearing is shown in Figure 6-15. It 

is a four-pad bearing with each pad being of 1.22 rad (70°) angular extent. A 

single pocket is incorporated in each pad. The pocket angular extent is 0.98 

rad (56°). Each pocket is individually fed through an orifice restrictor and 

two of the pockets have pressure tap? for measuring pocket pressure. The 1 

design clearance of the thrust bearing (per side) is 0.038 mm to 0.0432 mm 

(0.0015 to 0.0017 in.). The actual clearance is also 0.038 mm to 0.0432 mm 

(0.0015 to 0.0017 in.). 




D 




The thrust loads imposed on the seal test rig are shown in Figure 6-16. They 
are dtptndent on the pressure maintained in the outboard seal drain cavity, 
the bearing drain pressure and the turbine pressures. At design turbine 
conditions, the net thrust pushes the rotor toward the seal end and can go 
from 369 N (83 lb) with an outboard seal drain pressure of 517 kPa absolute (75 | I 
psia) to approximately 1432 N (322 lb) as the outboard dra.n pressure is 
lowered to 103 kPa absolute (15 psia). At low turbine power conditions, when 
its thrust is very Low, the net thrust would go from 309 N (69 lb) toward the 
turbine end to a net 529 N (119 lb) to the seal end as the outboard drain pres- 
sure is dec-eas^d from 517 kPa absolute to 103 kPa absolute (75 psia to 



6-S 



D 

D 
D 



15 p8ia)< Iti either case the net loads are well within the capability of the 
thrust bearing. 

Table 6-2 and 6'3 show thrust bearing performance at 7329 rad/s (70,000 r/min) 
and 10,470 rad/s (100,000 r/niin) for both liquid nitrogen and LOX. The tabu- 
lations show single side properties of the bearing and indicate variations In 
performance as a function of clearance. 

Load versus displacement for the total bearing is indicated in Figure 6-17. 
Because of light loading, the bearing will remain essentially centered with 
runner axial displacements less than 0.0025 mm (0.1 mils). At full displace- 
ment, the load capacity is approximately SllS N (1150 lb). Bearing flow for 
the double-acting thrust bearing is shown in Figure 6-18. In the centered 
position, the flow is 1.17 kg/s (16.2 gal/min at 70,000 r/min). Viscous power 
losses for a sinjiile-sided thrust surface are indicated in Figure 6-19. Pocket 
or recess pressure versus film thickness is shown in Figure 6-20. In the 
centered position, the recess pressure is 2585 kPa (375 psia), or half of the 
supply pressure. This pressure ratio provides near optimum stiffness. The 
double-sided stiffness is approximately 175 m-N/m (1,000,000 lb/in.). Figure 
6-21 is a plot of fluid-film temperature rise versus film thickness, for a 
single side. It is based on the assumption that all heat generated by vifcous 
friction is added to the fluid as it flows through the bearing and no heat is 
transferred through the bearing surfaces. In the centered position, the 
temperature rise is l.B'C (2.3''f) at 7329 rad/s (70,000 r/min). When the 
bearing clearance approaches 0.013 mm (0.0005 in.) the temperature rises 
dramatically and represents a failed condition. It is important to keep the 
temperature rise as low as possible to assure against vaporization in the 
bearing film. 

6.6 Rotor Dynamics 

The following three major analyses were performed on the 50-mm test seal 
rotor: 

1. Undamped critical speeds as a function of bearing stiffness 

2. Synchronous unbalance response 



6-9 



i. Stability analysi'J r.o e'uabli'ih synrrn stability, natural It rRiiuen- 
ciest and tnoda shapes. 

Because of t:he tarly emphasis on t:he spiral. -j^roove LOX seals, Che rotordynamic 
studies modeled the testier rotor as ouni'ittt»d for the LOX sp.al tests. Thf 
differences between the LOX seal rotor and the helium seal rotor are the mas-i 
and the polar and transverse moments of inertia of the seal runners. The LOX 
runner is greater in all three properties. Critical speeds or natural 
frequencies are higher and system stability is greater for the helium seal 
runner. Thus, if the analysis is not entirely accurate, it is at least 
conservative. 

In summary, the analyses show that s'^.al tests should operate satisfactorily 
over its entire speed range. The mofit troublesome natural frequency, the 
first bending mode, resides at approximately 13,611 rad/s, (130,000 r/min) and 
is well above the speed range. The lowest natural frequency, a rigid body 
conical mode, should occur at about 7329 rad/s (70,000 r/min) for the LOX 
r .or (higher fc<^ the helium rotor). The resulting vibrations should be well 
damped. Response analyses at 7329 rad/s (70,000 r/min) ;.ihow accaptable ampli- 
tudes for expected levels of residual unbalance, the stability parameters or 
grot/th factorj of all natural frequencies calculated are all negative and 
large, indicating good system stability. 

In practice, the dynamic characteristics of the tester proved considerably 
more troublesome than expected. Despite careful balancing and assembly, large 
shaft vibrations developed during both the rig checkout and the first three 
seal, tests, Vibration amplitudes started to grow around 4397 rad/s (42,000 
r/min) and reached unacceptably high shaft levels 0.025 to 0.038 mm peak to 
peak, (1.0 to 1.5 mils peak to peak) between 4816 and 5444 rad/? (46,000 and 
52,000 r/min). 

A brief experimental analysis revealed the following observations; 

1. Varying the bearing supply pressures between 3.79 and 6.20 MPa (550 
and 900 psig) did not eliminate the vibrations. 



6-10 



2, The nrbif", whilu np^.raiin^ M pressures nt 4,82, 5.51i ind b.20 MI'.i 
(700, 800, and 900 psig) were clo«e t.o (but not locked to) shatL 
speed. Orbits at pressures ot 3,79 and 4,14 MPa (550 and 600 psig) 
were exactly s/nchronuus. 

3. The speeds at which the vibration became Urge decreased somewhat as 
bearing supply pressure was increased. Also, the occurrence of 
noticeable bubbling around the seal end probe showed the same trend. 
Both effects are noted below: 



Supply Pre; 


ssure 


Vibration Threshold 


Bubbl 


ing Threshold 


(MPa) 




(rad/s) 


„„.,„..=™ 




(rad/s) 


6.20 




4837 






_ 


5.51 




4837 






3769 


4 c 82 




5026 






>3769 


4.14 




5403 






5151 


3,79 




54Q7 






5151 



4. Both the vibration and bubbling were most oevere on the seal end of 
the tester. 

5. Lower speed operation revealv.* small but noticeable synchronous 
vibration peeiks at 3141 rad/s (30,000 r/min) and second harmonic 
peaks at 2450 rad/s (23,400 r/min). The latter may correspond with a 
natural frequency of 4900 rad/s (46,800 r/min). 

6. Because of the bubbling occurring around Che shaft probes and Che 
nonsynchronous behavior at the higher pressures, the mode shapes 
could not be determined. 

From these analyses, it was concluded that the design stiffness and damping 
properties of the seal end journal bearing were not being achieved. This 
would have the general effect of lowering the natural frequencies and decreas- 
ing the damping. The following comparisons provide additional evidence of 
weakened bearing properties. The first and second undamped critical speeds 
are predicted to occur at 7329 and 9423 rad/s (70,000 and 90,000 r/min). 



6-11 



re'3pecti"ely based on bearing st i Hntjsi'jy's ui 117.3 m-N/m (6,7 x 10^ tb/in.). 
If, however, the stitfness degraded to about 17.5 m-N/m (1.0 x lO^lb/in.) the 
same analysis would predict 3141 rad/s (30,000 r/min) tor the first v".ritical 
speed and 4816 rad/s (46,000 r/min) tor the second. Both speeds coincide with 
measured peaks in vibration. 

The occrrence of self-excited vibration of frequencies very ijlose to the 
synchronous frequency is very unusual. However, additional analyses of the 
tester revealed that such an instability can occur with sufficiently degraded 
damping. 

The most likely cause of poor stiffness and damping in the seal end bearing is 
vaporization in the bearing film. Experimental observations provided ample 
evidence of gas bubbles when the vibration occurred. Previously completed 
design studies indicated that the LN2 bearing fluid is nk,ar saturation condi- 
tion as it exits the bearing. An additional factor is the windage loss 
resulting from the LN2 contacting the end of the seal runner. Thia generates 
a substantial amount of heat which would transmit to the surrounding parts and 
increase the tendency for vaporization. To eliminate the windage loss and 
hopefully the '/aporization, the second labyrinth seal was installed (see 
Section 2.0) prior to the fourth seal test. 

6.6.1 _ Rotor ModeJ. for the 50-nim LOX Seal Sha£t 

The rotordynamic models used for both the 50- and 20-mm LOX shafts are shown 
in Figure 5-22. Table 6-4 indicates pertinent information relative to tne 
various mass stations and shaft elements for the 50~mm LOX seal rotor. The 
model contained 20 maas stations with disk elements located at Stations 5, 9, 
and 16 representing the seal runner, thrust runner, and turbine wheel, respec- 
tively. 

6 .6.2 Undamped Critical Speeds for the 50 -mm LOX S^ 1^ Ro t or 

The undamped critical speed map for the 50-mm seal rotor is shown in Figure 
6-23. The bearing stiffness variable is the diagonal stiffnesses along the x- 



6-; 2 



i 



n> 



I 



t 



I 

I 

I 



i}r y-<ixes. For tne 50-nvn LUX -ieal ibuii, the beatriny; support ",t;i 1 Infa'i;; wn;j 
taken tu be 117.9 m-N/m (6.73 x 10"' lb/in,), reprerientati ve of a radial clear- 
ance of 0.019 mm (0.00075 in.). At an operating speed of S86i rad/s (56,n"0 
r/min), the rotor speed is below the first critical, which is a rigid body 
mode. The first bending critical, which is the third critical speed, occurs 
at approximately 13,611 rad/s (130,000 r/min) which is more than two times the 
cperating speed. 

6.6.3 Synchronous Unbalance Response of y}£,,50^nOTi_ LOX SejJ^^£tor 

In computing the unbalance for synchronous response studies, API balance spec- 
ifications were employed. This specification indicates what normal iccepta- 
ble and attainable unbalance levels are. 

The formula for computing the unbalance iis: 

Ux « 5 (56,347)(Wt)/N2 (6.1) 

at 7329 rad/s (70,000 r/min). For the 50-mm LOX seal rotor: 

Ux = 0.197 g-mm (2,7J x 10"*^ oz-in.) (6.2) 

at 10,470 rad/s (100,000 r/min). For the 20-mm LOX seal rotor: 

Ux = 0.082 g-mm (1.145 x lO"'''^ oz-in.) (6.3) 

Wt * Rotor weight (lb) 

The above numbers are low levels and difficult uo obtain. Therefore, MTI 
designed for an unbalance level of: 

Ux = 0.36 g-mm (5.0 x 10" oz-in.) 

The unbalances of this magnitude were applied at each of the Stations 5 and 
16, representing the seal runner and turbine wheel, respectively. Two sets of 
runs were made; one in vhich the unbalances were in-phase and one in wh^nh the 
unbalances were 180° out-of-phase. In-phase unbalance excites the lateral 



6-13 



rigid body modes. i)uL-'»t-phast' unbalatiKP excite-i the conical rigid body 
modes. 

Figure 6-24 shows haLf amplitude synchronous response ac Stations 8, 10, and 
n which are the seal end journal beiring, the turbine end journal bearing, 
and thrust bearing, respectively with in-phase unbalance. At the operating 
speed of 5863 to 6073 rad/s (56,000 to 58,000 r/min), displacements are very 
well controlled? amplitudes are approximately 0.00025A mm (O.Oi mils). The 
maximum speed of the rotor is 7329 rad/v (70,000 r/min) and wiil be run at this 
speed with the helium seals. Amplitudes at these stations at 7329 tad/r, 
(70,000 r/min) are still very gmall and pose no problems. Figure 6-25 shows 
half amplitude response at Stations 9, 10, and 15 which are the thrust bearing 
'Stations 9, 10) and in the vicinity ot the turbine wheel. Response at maxi- 
mum speed is less than 0.00127 mm (0.05 in,). Figiire 6-26 rihows response at 1, 
16, and 20. Station 1 is at the seal end extremity ot' the rotor; Station 16 
is at the turbine wheel and Station 20 is at the extreme turbine end at the 
rotor. At the operating speeds, amplitudes are approximately 0*00127 mm (0.05 
mils) at 5863 rad/s (56,000 r/min) and 0.00254 mm (0.1 mils) at 7329 rad/s 
(70,000 r/min). Figures 6-27, 6-28, and 6-29 show similar plot s for out-oJ 
phase unbalance. For the out-of~phase unbalance conditions, the maximun. 
amplitude is 0.00381 mm (0.15 mils) at Station 1 (see Figure 6-29). 

6j|_6^.4_ Stabi^Uty Analjr^siG f£r the 50-mm^LOX Seal Rot^)r 

As mentioned in the introductory paragraph, the stabii'ty analysis produces 
three significant items of information. These are a stability parameter or 
growth factor, natural frequencies, and mode shapes associated with the 
natural frequencies. Negative growth factors imply a stab'e system. 

A tabulation of modal growth factors, natural freque: ,xes and mode shapes is 
indicated in Table 6-5. The first mode is a rigid body mode. The natural 
frequency is 6384 rad/s (60,983 r/min) and the growth factor is -1139.2. The 
mode is very well damped. It colacldes with the bottom line of the undamped 
critical speed map shown in Figure 6~23. Note that theirs is significant 
motion at the bearing stations, so that bearing characteristics influence this 
particular mode. Mode 2 is again very well damped. The natural frequency of 



6-14 



I 

E 
I 
C 
I 
I 
I 

I 

I 



thi'i mode i". beyond thfi range ot i h»> maximum rot. or spoed. The third and t i J f h 
modes show signil'icanL bending. Mode 4 in predominantly a lateral ri>?,id body 
mode. 

6.7 Thermal Analysis 

Using computer codes available at, MTI , in-depth stxudies were made to esl:abli,sh 
the temperature distribution in the test rig. The original study waa 
performed for the 50-mm LOX seal configuration using LOX as the working fluid 
for both the bearings and seals. The results are shown in Figure 6-30. It wa'j 
assumed that the LUX entered the bearings and 'jeal compartment at -IJi'^C 
(-280"?), and that the nitrogen supply r.o the turb'.ne and buffer seal was at 
21°C (+70°F). Fluid temperatures are indicated In bold numbers. The princi- 
pal item of concern with respect to the fluid temperatures is that: the LOX 
remain in a liquid state in and near the bearings to prevent bearing problera;i 
and erroneous capacitance probe readings. Figure 6-31 indicates drain pres- 
sure requirements versus fluid temperature rise to keep the LOX completely 
liquid. 

The liquid at the journal bearing drains and the capacitance probes which are 
at "zhe inboard drain locations is relatively cool, -ITS'C (-279°F). The 
temperature rise above the normal boilini.? point is IQ^C (18°F). Thus a total 
backpressure of tw'j to three atmospheres should be sufficient to maintain the 
LOX discharges in the liquid state. The housing and shaft temperatures are 
all quite reasonable. The largest temperature differences occur in the heat 
dam and buffer seal regio'i, as would be expected. The heat dam is calculated 
to be affective in preventing transfer of he.at from the turbine end down the 
shaft. 



6-15 



D10 



I 



Test Seals 

and Runner 

(See Figs 2-1 

and 2-2 Details) 






E > 

6 S3 




Seal 
Housing 



RH Journal 

Thrust Bearing 

Housing 



D10 



Heiium Labyrinth 
Buffer Seal 

No2zle Box 



TurbtneWhMl 



HealOam 

25 - 30 mm 
Clearance 



Fig. b-l i:ri>bt. Sicl i«)ii i)l Icsl Ki|>;; Housing Set I ionii IdeiU i t itd 



■5114&-3 



_. L. 



H 



c^ r 





•< 3 



S«3dl HinlbKUJ 

|(l OX ^ >iibinj 
Showfi) 



I adynrttli 
Se-al tiotisKitj 



Kijj. b-J Test Rig housing!: 




FLg. 6-3 T«sc Rig Nozzle Box 



GKlClfJAL PAGE IS 
OE POOR QUALITY 



6-18 



LN2 Bearing/ 
Helium Seal Oram 



Helium Seal 
Outboard Drain -J» 



I 



vO 



Lr42 Supply 
Journal Bearings 



N2 Turbine /Helium Seal 
Oram (port not shown) 

H2 Turbine lr>iel 




LN2Beanng/ 
Helium Supply Helium Seal 
Test Seals mboard Oram 



Helium Supply 
Labyrinth Seal 



Fig. 6-A Test Rig Fluid Systeas 



Mn4»? 



I 

O 



A — Journal Bearing Inlet 

B — Pressure Tap 

C — Capacitance Probe 



Restrictor 
Element 
(Typical) 





ciriir;H i »^(- ) 



Drain 



A — Thrust Beanng Inlef 

B — Pressure Tap 

C — Vibrafkon Piston InM 



SECTION D-7 
Right-Harxl Bearing Housing 



SECTION D^ 
Right-HarKl Bearing Housirtg 



Si- 






J. 






Fig. 6-3 Supply and Drainage - Turbine-Side Bearing Housing 



131171 



rrirriiirrr'r^crtr^CDC!!! 



cr a ?r* 



¥1 



I 



B 






^.^ 




A — Thrust Bearing Inlel 
B — Pressure Tap 
C — Vibration Piston Inlet 
D — Capacitance Probe 
E — Journal Bearing Inlet 




(cr^ri" 



SE CTION A- 7 
Lett -Hand Bearing Housing 



A — Journal Bearing InM 
B — Prsssure Taps 




Drams 



SECTION A -3 
LeA-HarxJ Bearing Housing 



Fig. 6-6 Supply «nd Drainage - Seal-Side Bearing Housing 



I 

K> 



// 



\\ 




« 



[ 1 ^ A - Njfc 

*.• V OK.. 



Inlet 
B - N2 LOX Drain 
C - N2 Dfain 



SECTION D 10 
HelikJB Labyr inth Seal Housing 




Thrust and Journal Beanr«g Drams 



SECTION A-9 
Shun Ptale 



Figure 6-7 Supply .ind OrainaRe - l^hyrintli Seal ai»d Shi* Plate Housingis 



miio 



C (n 



I 






~~Ji3 kW 
-^"^ (18 hp) 

2 6 kW 
(3 5 hp) 




Assumed 
All Liquid 



H_ 



Windage 25 6 kW (34 4 hp) 

Bearings _48 kV. ( 64 hp) 

Tom 304 kW (403^1 



Fig. b-8 Test Rig Power Losa«:s 



4aa»i«2 



I 




X 



Design Point 

Speed 7329 rad/s (70.000 rpm) 

Tip Speed 284 m/s (865 ft/s) 

Inlet Pressure 1034 kPa. abs (150 psia) 

Maximum Acceleratior Rate 274 m/s' (900 ft/s') 

ij 07 

Ornax ?06 8 MPa (30,000 psi) 




•2M30-1 



Fig. h-9 Turbine Doslgii Schematic 



m Oj C3 



c: r:: cr o □ 



rr: cr: c c 



OWGINAL PAGE IS 
0£ £OOR gUAUTX 



. '^'" IF' !^ 







..^^ ' )<— .^ 



Pig. 6-10 AasetnbLed Test Rig Shaft 



6-25 



75 



I 



67 



45 



37 



100 1- 



^ 


u. 

x: 


zt 60 


•«-• 


•*-• 


3 


3 


Q. 


Q. 


■*-• 


■*-> 


D 


o 


o 




0) 




.E 52 


c 
jD 


^ 


1_ 


k. 


3 


3 


H 




Conditions 
Gas Ng at21='C(70°F) 

Fixed Pressure Ratio 10/1.0 
Speed 7329 rad/s (70,000 rpm) 

Mass Flow A Pi = 1034 kPa, abs (150 psia) 

is 0.40 kg/s (0.90 Ib/s) 
Mass Flow a P-i 



J 



200 250 300 

Inlet Pressure (psia) (P^) 



350 



400 



1034 



X 



1379 



1724 



2068 



2413 



Inlet Pressure (kPa, abs) 



2758 



B3793-I 



Fig. 6-11 Turbine Output Power versus Inlet Pressure 






< t 



I : :-=-'■= -* t ^ \ f**"*^ 



«s*aiii f,t~^.-\ » -. i • -i- » •» 



1 
to 



1.0 i- 



0.9 L- 



c 
'« 
O 



f^ 0.8 — 



0.7 



X 



Design Speed 



Conditions 
All Conditions Fixed Except 
rpm; i.e., Load, P1/P2. etc. 



I 



1 



1 



J I 



50,000 60,000 



70.000 80,000 
Speed (rpm) 



90.000 100.000 



J_ 



5,235 6,282 



7.329 8.376 

Speed (rad/s) 



i 



9,423 10.470 



Fig. 6-12 Turbine Efficiency Ratio versus Speed 



83792-1 



H ' I 



Jjij 



•6' 



MM4HHM »UU|P CONMHI* 



-MtMom MATi. rait »*i. j* ictMa 




Z.42S 



•t.vto 



aiwv.l>MC1. , 







KALI' ZO/I 



ORIGINAL PAGE IS 
QEPOOR QUALITY 




■°°| 0M> . e< QC««l>- 

'(II HOU OM *y gl5j 0C» D.C. 

|-»| B |,ooa5 i*| 



|>l6|.oo<!)^ 



.400 




i Mouts CO sr ON 

K N .OOP ^<N %C 



2Z ectota to « »c«o 

■♦IBI.0O4TOTA1.I 
(SK cxrM. E-1 BOC 
BLAOC PWIFILS) 



i£2LDiQm: ERAME 



/ fSiJ ULAl 




Ollk 






© 



.840. 



Z K 



TTi r 



Fig. 6-13 Turbine Wheel 



■ 



JL 



■»ntiu: UMC 



.UVtiOl*) 





u 

•moc raocikC v lu. aktuko. mamtain suabp 

WrjAC tCADI MW VAtI MIIUIN MAXIMUM AUO 
MINIMUM -OLeiAMCt BAKI09 tuT »WA|.L K A 
1M001U 'UN3ITION. 



otTAiL e-u fC-S) 

4CAi.li ZO/I 



^«» 



I 



ORIGINAL PAGE K 

QE POOR quauty: 




|-»|B|.OlM-fifMn 
(9H OCTAK. E-H IOC 
BLAOC PWWILC) 






o 



NOTeat 

I MAT«Q1AL • ALUMir*IM A'JJJY lOm-lt, • PiNCAKl WRaiNCi . K3IUilNI4 
'O 6e MAOt «0M &A» VOCK UAVlNd A MINIIuHJM i-SNalW Of 

tu«« ,Ji TiMe» PoeiiiivKj TuicKNesa awo aurFiCiENT oiamctBr 

to ACMItve ZCOUIKSS WUUMI 

I. uuRkMHic. 't^si-txa no«(i« M^cv^^Hmc■ »m m t \ %*ic » 

i.uf*isa oiHcawise spkibmo coNCENTdiciTf TO ac utLO wituin 
,OOiO"A "0 -6- 

*. wwnu TO M loiN TtaTEo at 110,000 rpm, MeAsyae ano okooo 
wutn. oj-raoe oiAwerBawio diameters -K- vh-b- mpobe auo 

AFTCIZ "tCV. NO CuAMAft PftONtOVOte. SIL HOTl-t 

&.OrNAMICAI.U^ SALAfoCe WMCBU tU FUANLS A ^% WITUIU .CQOSOtlf^. 
WMOwe MATEWW. AT ASBA S iWOICATBO. 

«,AMO0IEE N AC^aOAWCB WITH AM» ZWI MfV. liN\.M«.\n<i 

A»4o WIN •' ear 

X DATUM [TcTJl'. DCFiNOS BY THE ^VIfll^,^6E OP POINTS •MIDWAY 

afTWEE'J liLAOEb AT i.il? OIA, BASIC, 
a »WO^^. \ MTIR S»\H TtSI »V.OyM<.>*«T »t»0»TM.HT ms»vcT 

»l«. AMS tUAIi. \00-,. liUMKtt K«lS.HO lU»\tKHOHS »Wn\%S\«Ul. 
r.miPECTlON I^EPORTi A,NO MATC1^\^V.C.EB,T1F1CA.T10M WaHUIWLO. 



PRECEDING PAGiS tfLANK NOT FILMED 



joi*. 



2_ EOUIOUT FRAMH 



I TT 



tamm OMt meitw 



ss 



'— CT-J Mechgm ca l 'faohnobgy 

-— ^W-JB n i — M WW 






TuBOIloe wMceu 



MT41 I E I -Qoeoos 



IjacjOl 



3e: 



furbine Wheel 



6-29 



/SM 



. LOX Supply 



ON 
I 




C 



> 

O 
CO 

pi 

2: 

o 

s 



(180°) 

30 mm Dia. 
(1.1828 in.) 



A 

L 



0° 



jr/2rad 
(90°) 



TT rad 
(180°) 



3b'/2 rad 
(270°) 



2?r rad 
(360°) 



/ 
/ 
/ 

/ 



(72°) >Z 



^ 




4.72 mm 



I 



jr ^*1^"*~(6'.i86in!)[^| /^ K 

9//////////X//?7777777777Z//>^^ 



"ZZZW/. 



P 



y///////// 6\ 



T 



I 

' 
/ 

/ 



J 22.0 mm 
(0.866 in.) 



2.82 mm 
(0.111 in.» 



Developed Detail of Journal Bearing 




Section A-A 



Final Clearance = 0.0191 - 0.0254 mm (0.00075 - 0.0010 in.) 



Fig. 6-14 Test Rig Journal Bearing Recess Geometry 



1 



^ ^70.356 mm 
(0.014 in.) 



822812-1 



I 

N3 



LOX 
Supply 




— 0.51 mm 
(0 020 in.) 



Bearing 
Pad 




1.22 rad 
(70°) 



Section A A 



1.30 mm 
(0.051 in.) 



Fig. 6-15 Test Rig Thrust Bearing Recess Geometry 



822762 



pHB| P^A3*4 ^.^>,m*l fjAK.] r..^^'-'i f '---') f:,t>i-ri 

Im m-ju-^j B W. i wnii'i-^u * »iss: nr.T-r:* •=---"-=^;^ *cx^!qn=r;* m.z::^=^-~. i*==4:r= ^-4 



i_„ 



1-1 fki^i iNiiai^ t- 



I 




FfTfT 



738N 
- (166 lb) 



Pi = 517 kPa. abs (75 psia) F^ = 1490 N (335 lb) 
P2 = 517 kPa. abs (75 psia) F2 = 1125 M ( 253 lb) 

Net Thrust = 236 N (53 lb) 



Pi = 103 kPa. abs (15 psia) 
'' 517 kPa. abs (75 psia) 



Fi = 552 N (124 lb) 
F2=1156N(260lb) 



Net Thrust = 605 N (136 lb) 



Fig. 6-16 Test Rig Thrust Loading 



822815-1 



> ^. i . iii i ii i i«ii 






CO 
o 



5000 r 



4000 



3000 



2000 



1000 - 



0^ 





1200 




1000 


^ 


800 


CO 
O 


600 



400 



200 







10.470 rad/s 
(100,000 rpm) 




7.329 rad/s 
(70.000 rpm) 



J i__j I I I II I I I I I I I I I 

02 0.4 0.6 0.8 1.0 1.2 1.4 1.6 

Displacement (mil) 



I I I I I I I I I n I i I I L 



J I I I 







0.008 0.016 0.024 0.032 

DlsF>lacement (mm) 



0.040 



Fig. 5-17 ThruHL Bearing Load Caparity \'itsus Displacerafiil 



851585 



M KHK" Ijaai f^-*-* f""^ p^^ 1-^— *^ f^^ famm^ 



■^BSwBI ShHMK SnMMP^B 



1.2 r 



1.1 



«- 1.0 

I 0.9 
u. 



I 



0.8 



0.7 



E 
o 




10.470 rad/s 
(100.000 rpm) 



J L 



-I u 



J 1 1 1. 



0.2 0.4 0.6 0.8 1.0 

Displacement (mil) 



1.2 



1.4 



1.6 



I I I i__l I I I L 



J 1 ! I I I L 



J I 1 







0.008 



0.016 



0.024 



0.032 



0.040 



Displacement (mm) 



Fig. h-lH Tluii.-^t Ik'.iring Flow versus Displati-raeal 



851596 



I 



(O 

O 



o 
a. 



o*- 



CO 

O 



o 

Q. 







Bearing Data: 

Supply Pressure = 5.171 MPa 
OD = 58 mm ID = 32 mm 
Orifice Diameter = 2.09 mm at 7.329 rad/s 
= 1 .73 mm at 10,470 rad/s 
Discharge Coeff.= 0.9 



10.470 rad/s 
(100.000 rpm) 



7,329 rad/s 
(70.000 rpm) 



J- 



J. 



2 3 

Film Thickness (mil) 



u 




_l u 



J 1 1 1 



0.02 0.04 0.06 0.08 

Film Thickness (mm) 



0.10 



0.12 



Fig. t)-iy lluiisl ik-ariiig Powit Loh;. - SingU- Sidi- 



851597 



Mi p™.^-^ f--«-l 



f «' "» r «« I 



i.^..-i ^ ^^^4 fwMHi fanH^ p**"* 



* f •. i 



800r 



CO 
Q. 



(D 



CO 
Oi 



CO 
CO 
0) 

u 
o 
tr 



U5 



5.0 



4.0 



3.0 



2.0 



1.0- 



0^ 



CO 

Q. 

o 

^. 
3 
CO 
CO 
<D 



CO 
CO 

o 
tr 








Bearing Dala: 

Supply Pressure = 5.171 MPa 
OD = 58 mrn ID = 32 mm 
Orifice Diameter = 2.09 mm at 7.329 rad/s 
= 1.93 mm at 10,470 rad/s 
Discharge Coeff.= 0.9 



-10.470 rad/s & 7.329 rad/s 
(10C.OO0 rpm) (70,000 rpm) 



2 3 

Film Thickness (mil) 



0.02 0.04 0.06 0.08 0.10 

Film Thickness (mm) 



0.12 



Fig. b-IO I'lniist ik-aring Kui t'h^. Pn-risuri' 



851 598 



[^ 



-4^- «' ii .t.mi 



I 

00 



50 



40 



O 



i 30 

2 
<u 
a 

E 
I- 



20 



10 



O"- 



100 r 

90 
80 

u: 70 

o 



CO 

£ 

® 
a 

E 
a> 



60 



50 



^ 40 



30 
20 

10 








Bearing Data: 

Supply Pressure = 5.171 MPa 

OD = 58mm ID = 32 mm 

Orifice Diameter= 2.09 mm at 7.329 rad/s 

= 1.93 mm at 10,470 rad/s 
Discharge Coeff.= 0.9 



/- 7,329 rad/s 
(70,000 rpm) 




10,470 rad/s 
(100,000 rpm) 



X 



2 3 

Film Thickness (mil) 



0.02 0.04 0.06 0.08 

Film Thickness (mm) 



0.10 



0.12 



Fig. 6-21 Tlinisi Ik-ar iii}^ 'IVnipi.'rat iirt' Rist' vtTsu.s Film Thickiit-K.s 



851599 



cs E c ^ £3 a ■'^^'3 e i~ a ir: c: 



ON 
i 



79.2 mm- 




rJ? 



© 



D=-^ 



@ © 







Bearing 
No. 1 

9 



-104.6 mm 



50-mm Seal Shaft 



20-mm Seal Shaft 



® 



© Q 




® @ 



@@ 



Fig. 6-22 Rotordynamics Model of Test Rig 



851602-1 



CAOCHIC It 



I 

O 






t LOX S ALUniNUH UHEEI. t S« HH t 
CfflTXCAL SPEED NM 



R 


T 


II 



E 




P 4 



«< iST GMTICM. 
B> ■• CMITICM. 

c- an attricM. 



OPEMTINQ SPEEi 



KMINO STIFFNCSf (U/IN) 

Fig. 6-23 Undamped Critical Speed Map - 50-mm LOX Seal Shaft 






cacacstaE-^crseccsc! 



ON 
I 



ItTI 



• •• 



CADCNfE at 
WIAUWWt RfffPONf I 



t ALUniNUN UHCCL < IN PHAK « PLOT NO. I « 
MLF M1PLITUDS VI fPEIB AT lELCOTIB ITATIOHI 




•.I 



•TATION0 
A> I 
I- It 
0* li 



9 



tMMUMCCI 
COZ.IH.l 
AT fTATIOHt 



BU,« t.MC-«3 



• .4 



•.a 



iff ED (Rfll) "*•' 



•.I 



t.« 



1.1 



1.4 



Fig. 6-24 Half Amplitude Response, Stations 8, 10, 11; in-Phase Unbalance, C = 0.0191 



nun 



I 



HTI 



CADENIE li 
UNIALANCE REIPONIE 



« AlUHINUn UHEEL t IN PHASE f PLOT NO. • t 
HALF AHPLITUDE Mi IPEED AT SELECTED ITATIONi 




S.« 



• .I 



• .4 •.• Ii.t 

•PEED (BPt1> •*»•* 



t.t 



t.l 



fTATtOM 
A> I 
!• 3t 

e* IS 



? 



IMMUMCCf 
fSZ.M.) 
AT STATIONS 

SU^- •.s«c-«a 



1.4 



Fig. 6-25 50-inm Seal Half -Amplitude Response, Stations 9, 10, 15; In-Phase Unbalance, C = 0.0191 



OBI 






KeeEac3C5Ke^C3EKC3C3CC3C 



I 




N 



T 
u 

D 



N 
I 



HTI 



CADCHIC 11 
UNIAUNCC RCIPONIE 



« ALUniNUH UHEEL 1 III PHMC t PLOT NO. 3 » 
HALF AWLITUOE Ut tmEED AT lELCCTEB tTATIONi 




ITATIOM 
A- i 

e* M 



? 



UNMUWCCf 

lez.iH.) 

AT tTATIONt 



It u^* t.sw-Aa 



••• 



•.I 



•.4 



••• t.l 

•l»CCB (HPfl) '<*•' 



t.t 



1.1 



t.4 



Fig. 6-26 50-imn Seal Half-Amplitude Response, Stations 1, 16, 20; In-Phase Unbalance, C = 0.0191 sm 






WT! 



CADENSC It 
UNIALANCE RESMNiC 



t ALlMINUn UHEEL t OUT OF PHASE t HOT NO. i t 
HALF AWLITUDE Uf SFEEO AT SELECTED STATIOHS 




STATIONS 
A- i 
i> S* 
ۥ tt 



? 



WMUHCtt 
(02. IN.) 
M STATIOnC 



S U,* •.M6-«3 



SPEED (RPB) >**•* 



Fig. 6-27 50-imn Seal Half -Amplitude Response, Stations 8, 10, 11; Out-of-Phase Unbalance, C = 0.0191 



nun 



f _ I t X n - ^ i »- J »- » » ' !^SB| f^^ iar- 1 HI y^^ ■ - ' « J •— » » -i-! ii—- « y^^^ P^ 






HTI 



CADEHSE It 
UNIALANCC MSPONfZ 



« ALUniNUn UHEEL < OUr OF HMSE « PLOT NO. I t 
HALF AIVLITUDE VS fFEIED AT iELECTEO STATIONS 




STATIONS 
A* • 
i« IS 
C- IS 



B c 


w 


s 


•2 


o 


i 


1—1 

> 

r 


> 

r 


o 

t?5 


^ 


H^ 


>< 


W 



? 



inSAUMCCS 

cox.::;.} 

AT STATIOMS 



s u,* t.s«c-t> 

ts u-«-«.Mc-«a 



SreED <Rffl) "•* 



»e4 



Fig. 6-28 50-inin Seal Half-Amplitude Response, Stations 9, 10, 15; Out-of -Phase Unbalance, C = 0.0191 



mm 



0^ 
I 



NTI 



CADCNSE It 



t ALUniNUn UHCCL t OUT OF MMfC t PLOT t». 2 t 
HALF AnPLITUDC US iFflD AT tCLECTCD STATIONS 



A 
H 
P 
L 
I 
T 
U 
D 
E 



N 
I 




t.l 



••• 



STATIONS 
A« 1 
!■ IS 
C> M 



? 



UNMLMCCS 

f ez.iH. ) 

AT STATIONS 

S U.> •.S4C-93 
IS U,— •.SM-t3 



SPEED CRPfl) •*"* 



Fig. 6-29 50-mm Seal Half-Amplitude Response, Stations 1, 16, 20; Out-of-Phase Unbalance, C = 0.0191 mm 
pt^ |_,^| f^-:«« 9~-t r^l %^^i •_--! MM Ksa^ ph^J w---* *-~>''^ »-- -» »!--: = « fc-«i.< f^^ »s*sa mm^ foa 



TEMPERATURE DISTRIBUTION f C) 
6,280 rad/s, 50-mm SPIRAL GROOVE SEAL 



-173 -173 -173 



-156 



-148 



.-152. 



I 



-153 



-162 



^-155 = 



-154 



n 

-156 




r-159-. 
I ! 



-159— -162 



s 



n 



-157" 



4s 



I 



-157 



-173 -173 -164 



-1 73 r ~i 

-==L,_172 



158-154 



-172 



-172 



-34 




+21 



822824-t 



Fig. 6-30 LOX Seal TesL Rig 






10 
9 
8 
7 

? « 



LOX Seal Tester 




Saturation Temperature: 


-18yC(-29rF)at1 atm 


April 22, 1982 






Saturation Temperature Rise (° F) 



5 10 15 20 25 

Saturation Temperature Rise (° C) 

Kig. h-ll li ^,l Ki^ Drain Pressure versus S^itiiral iun Tc-Mpi-raUiri 



•51SM 



f) 



c C3 n a C3 C3 



I 

00 



E 

*-• 

CO 
CO 

<D 

1_ 

(L 

c 

'cO 

k. 

Q 



10 

9 
8 

7 
6 



LOX Seal Tester 

Saturation Temperature: -183° C (-297° F) at 1 atm 

April 22. 1982 




Saturation Temperature Rise (° F) 







10 15 20 25 

Saturation Temperature Rise {° C) 



Fig. h-JI li-si Kig Drain Pressure versus Sdluration lenperal uri- 



851534 



f^i,.« f^^-^l t, — I «■* •-' »» »-'• jp**l f--^'^ f» 



r«--:^-T » — ** » ' « -« «i-.-=ai *-•_!- « !»-!*»■»•? »— asai 



TABLE 6-1 
SUMMARY OF CONCENTRIC JOURNAL BEARING PERFORMANCE 



Liquid 


LOX 


LOX 


LOX 


LOX 


LN2 


LN2 


N, 3p««d, rad/i 


10,470 


7,329 


7,329 


10,470 


7,329 


10,470 


P(, Supply Prciiuri, kPa 


5,171 


5,171 


5,171 


5,171 


5,171 


5,171 


C, Radial Cltaranct, nm 


0.0191 


0.0191 


0.0254 


0.0254 


0.0254 


0.0254 


dot Orifict Diamctar, nm 


0.711 


0.762 


0.965 


0.965 


0.965 


0.965 


Pk, Racaaa Praiaura^ kPa 


2,620 


2,592 


2,613 


2,627 


2,551 


2,813 


q, Plow, kg/fl 


0.111 


0.132 


0.222 


0.191 


0.190 


0.180 


hp, Powar Lois, kW 


2.20 


0.82 


0.77 


2.07 


0.58 


1,56 


^xxt Stiffnaii, mN/m 


121.9 


117.9 


82.7 


86.3 


83.7 


87.0 


Kyy, 














Kxy, Crosa-Couplad Sti££na«^;, mN/m 


111.4 


64.6 


38.0 


64.1 


35.4 


48.5 


^yxt 














Dxx> Damping, mN-s/m 


21.2 


17.7 


10.3 


12.1 


9.6 


9.3 


Dyy, 














At, Tamparatura Risa, "C 


12.53 


3.93 


2.18 


6.82 


1.35 


3.82 


Mc, Critical Mass, kg 


4.45 


8.78 


6.17 


3.15 


9.98 


3.17 



6-49 



TABLE 6-2 



II 



SUMMARY OF THRUST BEARING PERFORMANCE 



(7329 rad/g; Orifice (do ■ 1.61 mm) 



m 



• I 













Pocket 


Clearance 


Force 


Flow 


Losses 


Temperature 


Prejsure 


(mm) 


(N) 


(kf^/s) 


(kW) 


('•O 


(kPa) 








LOX 






0.U4 


429.6 


0.786 


1.32 


1.03 


430 


0.0762 


915.0 


0.744 


1.43 


1.18 


910 


0.0666 


1155. 1 


0.724 


1.46 


1.24 


1144 


0.0572 


1490.1 


0.694 


1.51 


1.34 


1472 


0.0476 


1968.2 


0.650 


1.57 


1.48 


1932 


0.0381 


2659.2 


0.580 


1.64 


1.73 


2585 


0.0305 


3405.3 


0.497 


1.72 


2.12 


3272 


0.0229 


4301.2 


0.381 


1.83 


2.95 


4059 


0.0152 


5200,9 


0.228 


2.00 


5.39 


4772 


0.0076 


5821.8 


0.069 


2.36 


21.0 


5134 



1 1 



W2 



0.114 


414.9 


0.665 


0.98 


0.75 


418 


0.0762 


884.6 


0.632 


1.06 


0.85 


891 


0.0666 


1117.5 


0.614 


1.09 


0.90 


1114 


0.0572 


1444.5 


0.590 


1.12 


0.97 


1434 


0.0476 


1919.23 


0.552 


1.17 


1.07 


1893 


0.0381 


2603.7 


0.495 


1.22 


1.26 


2543 


0.0305 


3344.8 


0.425 


1.28 


1.53 


3228 


0.0229 


4245.5 


0.327 


1.37 


2.13 


4023 


0.0152 


5166.7 


0.197 


1.51 


3.88 


4754 


0.0076 


5808.0 


0.061 


1.79 


14.99 


5131 



IJ 



6-50 



^ : 



I 



I 



TABLU-3 
(10,470 rad/s; Orifice (dg * 1.55 nm) 













Pocket 


CleArance 


Force 


Flow 


Losses 


Temperature 


Pressure 


(mm) 


JNL 


(k^M 


(kW) 


CO 


(kPa) 








LOX 






0.U4 


425.2 


0.727 


3.6i 


3.04 


417 


0.0762 


914.2 


0.690 


3.89 


3.46 


8<)1 


0.0666 


1158.7 


0.671 


3.99 


3.65 


1126 


0.0572 


1502.8 


0.642 


4.10 


3.92 


1455 


0.0476 


1996.6 


0.601 


4.25 


4.35 


1922 


0.0381 


2709.2 


0.536 


4.44 


5.09 


2585 


0.0305 


3479.2 


0.458 


4.65 


6.24 


3285 


0.0229 


4404.0 


0.347 


4.93 


8.74 


4090 


n.Q152 


5307.6 


0.202 


5.38 


16.40 


4805 


0.0076 


5868.9 


0.057 


6.30 


68.25 


5142 


0.114 


406.3 


0.615 


2.67 


2.20 


402 


0.0762 


875.9 


0.585 


2.88 


2.50 


859 


0.0666 


1111.3 


0.569 


2.96 


2.64 


1087 


0.0572 


1443.5 


0.547 


3.05 


2.83 


1406 


0.0476 


1922.6 


0.51'. 


3.16 


3.13 


1862 


0.0381 


2619.2 


0.4^/ 


3.31 


3.66 


2514 


0.0305 


3380.9 


0.394 


3.47 


4.46 


3211 


0.0229 


4312.1 


0.301 


3.69 


6.20 


4026 


0.0152 


5248.9 


0.178 


4.04 


11.52 


4773 


0.0076 


5855.8 


0.051 


4.77 


47.81 


5138 



6-51 



TABLE 6-4 
50-a:M LOX SEAL ROTOR PARAMETERS 



NMAT 
I 



NGYRO 
1 



SO 



lUNITS 




ISCAL 




I PR EC 




MAT YOUNGS MOO. DENSITY SHEAR MOO. 

NC. |La/IN««2l lt.B/IN»93| (LB/IN»«2i 

1 3.t3000*07 2.960D-01 9. 40000^06 



ROTOR DATA 
























STAT MASS 




IP 




IT 


LENGTH STIFF. 


MASS 


INNER 


YCHJNGS MOO. 


DENSITY SHEAR MOD. 


NO. 


. ILBSI 




ILB-IN^OZI 


|UB-IN$»2) 


(iNi 


OiA. 


OiA. 


OIA. 


1I.B/IN««2} {LB/IN«^»3} J 


[1.8/IN»«21 


I 


0.0 




0.0 




0.0 




0.20L 


0.314 


0.313 


0.0 


3.13000»07 


2.960D-0 1 


9.4000*06 


2 


0.0 




0.0 




0.0 




0.570 


0.343 


0.650 


0.0 


.3; ^300O*07 


2.9600-01 


9.4000*06 


3 


0.0 




0.0 




0.0 




0.2S0 


0.500 


0.500 


0.0 


■ 3<:l3O0O*-07 


2.960O-01 


9.4000*06 


4 


S.200O- 


02 


9.900OO- 


■03 


5.50000- 


-03 


0.190 


.0^500 


0.500 


0.0 


3.1300D*^07 


2.960O-01 


9.4000*06 


1 


0.0 




0.0 




0.0 




0.500 


0«S90 


3^120 


0.0 


3.130QO«07 


2.9600-01 


9.4000*06 


6 


0.0 




0.0 




0.0 




0.875 


i«iao 


1 . 1 ao 


0.0 


3.13000«07 


2.9600-01 


9.4000*06 


T 7 


0.0 




0.0 




0.0 




0.437 


1 .1-80 


1.I60 


0.0 


3.1300D«07 


2.9600-01 


9.4000*06 


t5 8 


o.o 




0.0 




0.0 




l«6SO 


1.180 


1. ISO 


0.0 


3.130aO«'07 


2.960O-01 


9.4000*06 


9 


0.0 




0.0 




O.O 




-O.&OO 


1.180 


2.400 


0.0 


3.130004^07 


2.9600-01 


9.4000*06 


10 


0.0 




0.0 




0.0 




i«a75 


.1.1-80 


.1.180 


0.0 


3.t3000»07 


2.9600-01 


9.4000*06 


IB 


0.0 




0.0 




0.0 




0«£i4.0 


1 .1 80 


L.iao 


O.O 


3.130Q0*07 


2.9600-01 


9.4000*06 


12 


0.0 




0.0 




0.0 




OsSOO 


ti.^60 


1..060 


0.0 


3.13000*07 


2.960D-C1 


9.4000*06 


13 


0.0 




0.0 




0.0 




o..t.io 


0.670 


0.670 


0.0 


3.13000*07 


2.9600-01 


9.4000*06 


14 


0.0 




0.0 




0.0 




0.380 


0.670 


0.670 


0.0 


3. 13000*07 


2. 9600-01 


9.4000*06 


15 


1.0200- 


01 


2.20000- 


-02 


1.80000- 


-02 


0.490 


0.670 


0.300 


0.0 


3.13000*07 


2. 9600-01 


9.4000*06 


16 


1 .8700- 


01 


1.61000' 


-01 


8.60000- 


-02 


0.7S0 


0.670 


0.300 


0.0 


3.13000*07 


2. 9600-01 


9.4000*06 


17 


0.0 




0.0 




0.0 




0.450 


0.670 


0.670 


0.0 


3.13000*07 


2.960O-Q1 


9.4000*06 


18 


0.0 




0.0 




0.0 




0.150 


0.375 


0.375 


0.0 


3.13000*07 


2.9600-01 


9.4000*06 


19 


0.0 




0.0 




0.0 




0.190 


0.281 


0.28 1 


0.0 


3.13000*07 


2. 9600-01 


9.4000*06 


20 


0.0 




0.0 




0.0 




0.0 


0.0 


0.0 


0.0 


3.13000*07 


2.9600-01 


9.4000*06 


BEARING STATIONS 

e 11 












* 














30 MM SHAFT, 70900 RPH. C = .75 MILS, ALUMINUM TURBINE WHEEL 



» ROTOR PARAMETERS « 

HEIGHT OF SHAFT 

HEIGHT OF DISCS 

HEIGHT OF ROTOR, 

SHAFT LENGTH 

LOCATION OF C.G 



2. 976111 

1.77200 

•1. 718111 

10.62500 

1.37356 



Seal Runner 

Ip = 1.488 Ib-ln^ 

I^ = 2.125 Ib-ln^ 






f %'^1 



Thrust Runner 

' 2 

Ip = 0.401 Ib-ln 

2 
I(. = 0.201 Ib-in 



f '•"»** 



L2 



flSiiMMHi It^StSSi 



I 
I 

I 



TABLE 6-5 
50-MM LOX SEAL STABILITY ANALYSIS 



Mode 



Growth 
Factor 


Frequency 
rad/s (RPM) 


-1139 


6,385 
(60,983) 


-2247 


7,966 
(76,082) 


- 980 


8,472 
(80,921) 


-4863 


8,978 
(85,748) 


-1028 


9,995 
(95,459) 



Shape 




Comments 



Conical Rigid Body 



^ 



Some Bending 



Lateral Rigid Body 



Bending Mode 



6-53 



7.0 TEST FACILITY 

The helium seal testing was performed at Wyle Laboratory in Norco, California. 
Figures 7-1*, 7-2, and 7-3, provide several views of the test facility while 
Figures 7-4 and 7-5 are close-ups showing the tester installation. An overall 
view of the control room is shown in Figur" 7-6. The system control panel is 
depicted in Figure 7-7. 

The sections that follow provide a detailed description of the test facility 
including: 

• Fluid Supply Systems 

• Facility Controls 

• Instrumentation. 

7.1 Fluid Supply Systems 

Figure 7-8 is a simplified piping schematic of the fluid supply systems. The 
three major fluid systems supply gaseous helium to the buffer seals, liquid 
nitrogen to the tester bearings, and gaseous nitrogen to the tester drive 
turbine. Two additional systems supply gaseous helium to labyrinth buffer 
seals and gaseous nitrogen purges to the major fluid systems at various 
points. 

7.1.1 Helium Seal Supply System 

Gaseous helium is supplied to the facility from a 1550 m (55,000 scfm) tank 
trailer at pressures ranging from 3.45 to 17.24 MPa (500 to 2500 psig). A line 
from the tank trailer connects to a tank located near the tester. 

The helium flow to the tester then passes through a 10 micron (nominal) line 
filter and then through a dome loader regulator valve (40). This reduces and 
controls the pressure of the helium to the range of to 1724 kPa absolute (0 



*Figures are presented consecutively, beginning on page 7-15. 



7-1 



to 250 psia). Control pressure to the dome Loader is adjusted using a motor- 
ized regulator remotely operated from the control room. 

The '"low then passes through a calibrated Venturi flowmeter. Pressure (P5) 
and temperature (T5) are measured at the inlet to the Venturi as is the pres- 
sure drop from the inlet to the throat of the Venturi (P20). These allow accu- 
rate measurement of the helium supply flow. The helium then goes to a 
manifold near the tester where the seal supply pressure (P12) is measured. 



The LN2 originates at an 41.6 m^ (11,000 gal) storage tank near the test site. 
Transfer pumps at the storage tank supply LN2 through a vacuum jacketed line 






J 



The flow then splits into two lines and enters the tester (C9 A & B). In the i 1 
tester, the flow again splits with some of the helium passing through the 
inboard seal ring and some through the outboard seal ring. The former joins 
with the bearing drain flow. The latter enters the outboard seal drain cavity 
and passes out through drain ports (CIO A, B, C). 



11 



Pressure (P2) and temperature (T2) are measured in the drain line providing _, , 

both seal drain conditions and inlet conditions for a second Venturi flowmeter Hj 

located immediately downstream. The pressure drop across the Venturi (P18) is 

measured providing a record of the flow throi'gh the outboard seal. The || 

inboard seal flow is derived and is equal to the supply flow minus the 

outboard seal flow. | i 







The helium drain flow then passes through a pneumatically actuated pressure 
control valve (8) and is vented to atmosphere through a standpipe. The 
control valve is used to set the drain pressure between and 517 kPa absolute ^-. 
(75 psia), the latter being the nominal pressure of the bearing drain and J| 
hence the inboard helium seal. The valve is controlled either by a differen- 
tial controller using the bp.aring drain pressure (P3) as a reference or by j I 
directly controlling the pneumatic actuator with a remotely located hand regu- 
lator. The latter proved to be a more convenient method of incrementally 7 I 

iJi 
reducing drain pressure during steady-state runs. 

7.1.2 LN2 Bearing Supply System ''X-M 






11 ^ 



1-1 



to a 1.67 m'' (A40 gaL) elevated run tank at the test site. The tank is 
equipped with an automatic level controller which operates the tank fill valve 
(17) and keeps the tank about 3/4 full during operation. A 76-mm (3-in.) vent 
maintains the run tank at near atmospheric pressure. 

A line from the bottom of the tank goes to the pump fill valve (15) which 
connects to the suction of a single stage boost pump. The discharge of the 
boost pump connects both to the suction of the high pressure vari-drive pump 
and a motorized bypass valve (16) leading back to the run tank. 

The vari-drive pump provides the high pressure LN2 needed for operation of the 
hydrostatic thrust and journal bearings in the tester. It is motor driven 
through a variable speed mechanical transmission capable of operation from 314 
to 1152 rad/s (3,000 to ll,000r/min) . The pump is a two-stage centrifugal 
type capable of delivering 3.0 kg/s (60 gal/min) of LN2 at pressures up to 
6.894 MPa (1000 psig). 

The output of the vari-drive pump flows through a check valve and a 10 micron 
(nominal) filter into the LN2 manifold. The line is equipped with a burst 
disk designed to rupture at a pressure of 7.58 MPa (1100 psig) to prevent 
accidental overpressuring of the manifold. 

Two air-actuated control valves (5 and 6) connect to the manifold and supply 
LN2 to the hydrostatic thrust and journal bearings, respectively. These are 
shown in Figure 7-2. Each valve is equipped with a controller which automat- 
ically maintains a bearing supply pressure of 4.14 MPa (600 psig). A third 
air-actuated control valve (2) is installed in the LN2 supply manifold and 
provides a bypass back to the run tank. It is controlled by a remotely oper- 
ated, motorized regulator and set in conjunction with the speed of the 
vari-drive pump to provide a flow sufficient for the bearings. 

Each of the two bearing supply valves is connected to a supply manifold next 
to the tester. These are shown in Figures 7-4 and 7-5. Eight lines connect 
each manifold with the corresponding supply ports on the tester: C3 A, B, C, D 
and C8 A, B, C, D for the journal bearings and C5 A, B, C, D and C6 A, B, C, D 



7-3 



for the thrust bearings. Each of the 16 lines is equipped with a 60-mi«:.ron 
line filter to prevent bearing contamination. 



The tester is equipped with three groups of bearing drains. These are shown 

m Figure 7-3. Lines from five thrust bearing drain ports (C13 A, B, C, D, E) 

ant! three seal end journal bearing drain ports (C14 A, B, C) connect to a 

common manifold. Drain pressure (P3) and temperature (T3) are then measured 

in the line which is connected to an air-actuated drain valve (10). The valve 

connects to a common return line to the run tank. The pressure upstream of the ^ 

valve is automatically controlled 'o about 68.9 kPa (10 psi) above the satu- j|j 

ration presture of the fluid to preveit vaporization in the tester. 



7.1.3 GN2 Turbine Supply System 





11 



ll 



The third drain (Cl5 A, B, C) collects a mixture of LN2 from the turbine-side 

journal bearing and the helium from the adjacent labyrinth buffer seal. The 

line is instrumenr.ed for pressure (P4) and temperature (T4) and connected to 

an air-actuated droin valve (11). As with the other drain line, pressure is i<f^ 

(maintained at about (.'8.9 kPa (10 psi) above saturation pressure. ill 



tl 



li 



High-pressure GN2 is storod in a battery of tank trailers near the test site. 

A line from the trailers goes to a dome loader regulator (39) which reduces 

the GN2 pressure to the range of 1724 to 2068 kPa (250 to 300 psig). A run 

tank downstream of the regula'tor provides a local capacity. This is followed 

by a 10-micron filter. The Gl!2 flow then passes through two control valves 

'12 and 14), splits and enters the turbine through two diametrically opposed ^iJ 

ports (CI A, B). The turbine exhausts to atmosphere through a short tail 

piece. 






1 



y 



The first valve (12) throttles the flow to maintain desired speed. It is U. ;' 



•*•■ 



pneumatically actuated and can be controlled in manual or automatic mode. The 

second valve (14) is an emergency shutdown valve. It is a fast-acting air ^ I 

I*-': r 

solenoid type requiring air pressure to open. It is equipped with a spring ^'* 
which closes the valve on loss of air. A.'r pressure is supplied by a remotely ^ ^ 
operated electric solenoid valve designed to close on loss of power. 



ii 



7-4 



The emergency shutdown valve (14) is used to perform the high acceleration 
rate testing. To make a fast start, the speed control is switched to the manu- 
al mode and the diaphragm pressure of the speed control valve (12) adjusted to 
provide the correct start-up flow. The actual setting is determined by trial 
and error. Both tester overspeed trips are then lowered to slightly under the 
maximum speed desired for the acceleration run. The emergency shutdown valve 
is then manually energized. The valve quickly opens and the in-rush of gas 
rapidly accelerates the turbine. Upon reaching the maximum speed, the valve 
is closed by the overspeed trip and the tester coasts down. 

7.1.4 Helium Supply to Labyrinth Seal 

The labyrinth seals are supplied by the same tank trailer that supplies the 
helium seals. A separate line connects to the tank located near the tester. 
The line goes to a 10-micron filter, then to dome loader regulator (41). The 
downstream pressure is set to provide a 69 to 138 kPa (10 to 20 psi) pressure 
difference between the labyrinth supply and the adjacent journal bearing 
drain. This results in a typical labyrinth supply pressure of 517 to 586 kPa 
absolute (75 to 85 psia). A pressure sensor (Pll) is installed in the line 
which connects to the tester at C16. Inside the tester, part of the flow goes 
toward the adjacent journal bearing and mixes with the LN2 bearing flow. The 
combined fluids drain out through ports C15 A, B, C. (See Section 7,1.2.) The 
other part goes toward the turbine, joins with a small leakage flow from the 
turbine and exits the tester at ports C17 A, B, C, D. The connecting drain 
lines join at a manifold and vent to atmosphere. 

The original design of the tester used GN2 as the buffer gas for the labyrinth 
seal. It was found during checkout, however, that some of the GN2 was 
condensing and forming puddles of LN2 which ran into the turbine. The problem 
was solved by changing the buffer gas to helium which has a much lower conden- 
sation temperature. 

7.1.5 GN2 Purge 

The facility contains four GN2 purge lines These were used before and after 
all tests to prevent moisture contamination. Purge gas is supplied by a 



7-5 



7.2 Controls 



tJ 




862 kPa (125 paia) GN2 inatrument manifold supplied by Che high-pressure GN2 

tank trailers through a reducing regulator. The manifold is filtered with a 

10-micron (nominal) filter. Each purge line is equipped with a regulator 

followed by a remotely operated electric solenoid valve and a check valve. 

This allows individual flow rates to be set and independent operation of each 

line. One purge line goes to the helium seal drain cavity and a second to the 

labyrinth seal supply line. A third line is connected to the LN2 supply mani- 5 j 

fold while a fourth goes to the turbine inlet. The latter proved to be very 

important because of the open exhaust on the turbine. 









Due to the hazardous nature of the testing (especially in the LOX mode), all 
fluid components requiring on/off operation or adjustment during operation of | 1 
the facility are designed for remote or automatic control. In addition, vari- 
ous automatic shutdowns are incorporated in the controls. *" ", 

il 

Remote control is accomplished by on/off operation of a series of 12-V dc » , 

relays. The switches which energize the relays are in a panel in the control |J 
room which is shown in Figure 7-7. The relays reside in a terminal box at the 

test site and energize 110-V ac circuits. It v 







M; i> 



All four purge valves (19, 20, 21, 22) and the drain controls air supply valve 

(18) are electric solenoid valves directly controlled by 110-V ac circuits. 

The pump fill (15), tank fill (17) and turbine trip (14) valves are air-oper- if 

ated solenoid valves. The air to operate the air solenoids is controlled by 

electric solenoid valves energized by the 110-V ac relays. For the boost pump ^ „ 

and vari-drive pump, the 110-V ac relays operate 440-V ac motor starting 1 j 

circuits . 

11 

Control of the adjustable position valves and vari-drive speed control is 

accomplished using double-throw, momentary-contact, center-off switches. 1 I 

With the switch thrown in one direction, one of a pair of 12-V dc relays is " 

energized. Thrown in the opposite direction the other relay is energized. r -^ 

One relay causes a gear motor to slowly rotate in one direction increasing a ».J 
valve or motor speed setting while tht other relay causes reverse rotation of 



li 

7-6 f ] 



the same gear motor, decreasing the setting. For the vari-drive pump speed 
control is impLemented by the gear motor changing belt pulley ratio in the 
transmission. The pump bypass valve (16) is adjusted by direct rotation of a 
ball valve. The LN2 manifold bypass valve (2), helium seal supply regulator 
(25), labyrinth seal supply regulator (24), and turbine speed control (manual 
mode only) valve (12) are air-actuated devices whose settings are controlled 
indirectly by motorized regulators, themselves operated in the manner 
described. 

Several emergency shutdowns are incorporated. Emergency shutdown of the test- 
er is achieved by de-energizing the turbine trip solenoid. This can be done 
manually by a switch on the control panel or automatically by alarm modules 
operating on selected parameters. These include: 

• Overspeed (SI) 

• Overspeed (S2) 

• High outboard seal surface temperature (T12) 

• High inboard seal surface temperature (T13) 

• High tester housing vibration - axial (XI) 

• High tester housing vibration - radial (X2) 

• Low thrust bearing supply pressure (P14) 

• Low journal bearing supply pressure (P13) 

Each alarm contains a normally closed contact in series with the 12-V dc relay 
controlling the turbine trip valve (14). If any limit is exceeded, the 
turbine is tripped. All alarms operate in a latching mode requiring a manual 
reset before the trip valve can be re-energized. 

An emergency facility shutdown circuit and LN2 deluge system are also built 
in. They are controlled by separate switches on the control panel and are 
designed as safeguards to be used in the event of a fire or similar catastro- 
phy during LOX testing. Opening the emergency shutdown switch shuts off the 
power to many of the 12-V dc relay circuits. This trips the turbine; shuts 
down both pumps; closes both the LW2 tank fill and pump fill valves; and opens 
all LN2 drain valves. Opening the LN2 deluge switch opens an air solenoid 
valve admitting 172 to 207 kPa (25 to 30 psig) LN2 to a perforated manifold 



7-7 



adjacent to the tester. A massive flow of LN2 inundates the tester area and 
would help to smother an oxygen fire. 

Three automatic control systems are incorporated in the test facility. One 
controls turbine speed while the othar two are identical and control the pres- 
sure in each of two bearing drain lines. 

The turbine speed control system allows either manual or automatic operation. 
In the manual mode, the pressure in the diaphragm of the speed control valve 
(12) and, hence, the GH2 flow and turbine speed, is set using a motorized 
regulator operated from the control panel. This mode was used for all slow 
speed starts and was found to be quite satisfactory at high speeds. 

In the automatic mode, the valve is controlled by Woodward Model 2301 elec- 
tronic governor operating i'rt conjunction with a valve positioner. The desired 
speed is set using a potentiometer connected to the governor. A speed signal 
(SI) from the tester is electronically compared with the speed setting. The 
difference or error signal ia then amplified and sent fo the positioner which 
changes the diaphragm ptessure to raise or lower the tester speed to equal the 
desired setting. 

The LN2 drain pressure control system is used to automatically control the 
pressure upstream of the drain valve to about 69 kPa (10 psia) greater than 
the saturation pressure. A vapor bulb in the drain line is charged with 
nitrogen gas prior to cold testing. The charge pressure is calculated such 
that the fluid in the bulb is a mixture of gas and liquid throughout the entire 
range of drain temperatures during cold operation. Since the bulb and the 
drain fluids are at a common temperature and the bulb fluid is two-phase, the 
bulb pressure is also the saturation pressure of the drain fluid. The bulb 
pressure and the actual drain pressures are then measured by a differential 
transducer. The output signal of the transducer is fed back to the differen- 
tial controller which compares it with the desired 69 kPa (10 psia) pressure 
difference and adjusts the setting of the drain control valve to raise or 
lower the drain pressure accordingly. 



I* 



7-8 



I 



I 
I 
i 
\ 
\ 
\ 
I 

i 
f 

I 



7.3 Inatrumtntition 

Instrumantation was installed to measure film thickness, shaft displacement, 
shaft speed, tester vibration and various temperatures, pressures, and flows. 
Table 7-1* provides measurement and sensor details. The identification code 
indicates the type of sensor and its number; e.g., P5 is Pressure Number 5. 
The location code indicates whether the sensor is in the test rig (T) or 
installed externally (E). Those installed in the test rig are shown in Figure 
7-9 while the externa] sensors are depicted in the piping schematic shown in 
Figure 7-8. The test code indicates the tests during which the sensors were 
used; e.g., A - all, 1 - seal test No. 1, etc. The following sections discuss 
the measurement and sensor types. 

7.3.1 Film Thickness and Shaft Displacement 

Special design capacitance probes supported byMTI's Accv.measure'" System 1000 
amplifying and conditioning components were used for these measurements. 
Specifications for the Accumeasure System 1000 are given in Table 7-2. 



The basic construction features of the capacitance probes are shown in Figure 
7-10. Due to the large temperature range and aucicipated future use in liquid 
oxygen, special materials are used in their construction. The seal film 
thickness probes are constructed using 42% iron-nickel. This was chosen 
because its expansion rate is very close to that of the carbon-graphite now 
used in the helium seal rings. The shaft displacement prCfbes are fabricated 
from 304 stainless steel which provides good low temperature properties and a 
suitable expansion rate. 

Eccobond 104 epoxy manufactured by Emerson Cummings is used to bond the inter- 
nal parts together. The proper choice of adhesives was of great concern 
because of the need for good dimensional stability and high bond strength when 
subjected to repeated thermal cycling. Eccobond 104 proved to be considerably 
stronger and harder than several other epoxies and polyurethane adhesives that 



*Table3 are presented consecutively, beginning on page 7-30. 



7-9 



7.3.2 Seal Film Thickness Probes 



IJ 



were screened during Che design stage. Additionally, ic was found co retain |.J 

most of its bond strength when used in a typical probe construction and 

repeatedly cycled between ai'C and -196*C (yO^P and -321°F). | j 

While no LOX was used during the helium seal testing, the probes were designed | ] 

with its use in mind for future work. The principal hazard is material 

ignition. The metal parts are quite safe due to their high ignition temper- -r < 

atures and high thermal conductivity. The epoxy, however, is highly combusti- * I 

ole and definitely an area of vulnerability. To minimize this hazard, . , 

Refset", a special fluoroelastomer compounded by Raybestos Manhattan Corpo- j j 

ration, is used as a protective overcoating, both in the interelectrode 

grooves at the top of the probes and near the connector on the other end. The 

material was developed specifically to be used as a protective barrier and 

will not burn in an oxygen atmosphere. 



LI 













f. 



'^ I 



Two methods were used to measure seal film thickness during the testing. One 
uses a differential approach while the other provides a direct measurement. 
The former employs two probes for each measurement, one observing the back of 
the seal ring and the other the surface of the seal runner. The film thickness | 
is derived by measuring the difference between the outputs. Both vertical and 
horizontal measurements are taken on both the inboard and outboard seal rings. 
Two common seal runner probes are used, one serving both the inboard and 
outboard seal probes in the vertical direction and the other providing a simi- 
lar arrangement in the horizontal orientation. Figure 7-11 shows the differ- « li 
ential probe installation. 

The probe has an outer tip diameter of 5.0 mm (0.197 in.) and an interelec- 

trode diameter of 2.67 mm (0.105 in.). It is calibrated to have a range of ; J 

0.127 mm (0.005 in^) in air or helium. Figure 7-12 shows the detailed 

construction which followed MTl's standard practice and used the materials H ^ 

previously discussed. "» '' 

^ 






7-10 



The differencial method, which was employed during the first three seal testa, 
proved troubleaoma and yielded poorer results than we had hoped for. Several 
of the problems encountered are described as follows: 

^* Thermal Dij^tort ion_j)f_the Runner . With LN2 surrounding the inboard 
face and gaseous helium on the outboard side, the runner assumes the 
shape of a truncated cone. Thus with the reference probe measuring 
at the center rather than under the seal rings, the inboard seal film 
thickness would be understated while the outboard seal film thick- 
ness would be overstated. In either case, the amount is estimated to 
be about 0.0056 to 0.0085 mm (0.0002 to 0.0003 in.). 

2. Thermal Distortion of the Seal Housing. This is due to the temper- 
ature gradient across the seal housing resulting in the probes which 
are located closer to the bearings; i.e., the cold section of the 
tester, moving closer to their target than the more outboard probes. 
This has an opposite effect to that caused by the runner and tends to 
overstate the inboard film thickness and understate the outboard 
film thickness. The amount is a direct function of the temperature 
difference between probes (seal probes and runner probes) and is 
about 0.00183 mm/'C (0.00004 in./°F). 

A direct measurement of seal film thickness using embedded probes in the outer 
seal ring was made during the last seal test. Refer to Figure 2-23 for the 
arrangement and to Figure 2-2A for a photo of the actual ring. The tip geom- 
etry provides for a range of 0.076 mm (0.003 in.) The lead-off cables consist 
of 0.33 mm (0.013 in) diameter copper-clad cable chosen to minimize the 
mechanical impedance imposed on the seal rang. The diametrally opposed probe 
pairs allow measurement not only of the insuantaneouo film thickness but also 
the mean film thickness and the ring eccentricity. The 1.22 to 1.92 rad (70° 
to 110°) angular spacing of the probes is needed to avoid drilling through one 
of the three antirotation slots in the seal ring and to place the probe tips in 
the bearing pockets. The metal parts of the probe are constructed of 42% 
iron-nickel alloy to match the expansion rate of t'le seal ring. 



7-11 



7 , 3 1 3 Shaft D l3p I a e etne n t P ro be n 



7.3.4 Shaft Speed 



il 



Five probes are installed to observe the shaft, four radial and one axial. | I 
The radial shaft probes are installed in x-y pairs slightly inboard of the 

hydrostatic journal bearings. The probes are mounted in extension tubes and "* a 

secured to and adjustable from the outside of the tester. The axial probe is * ^ 

installed in the seal cavity and observes the edge of the helium seal runner -. . 

as shown m Figure 7-U. , I 

The radial probes have an outer diameter of 5.00 mm (0.197 in.) while the -f 

• k 

axial probe has a diameter of 3.18 mm (0,125 in.). The latter is conscructed 
without a grounded outer shield and has a smaller inner electrode to accommo- 
date a space restriction. Both sizes of probes are calibrated to have a range 

of 0.12/ mm (0.005 in.) in air and 0.191 mm (0,0075 in.) in LN2. Figure 7-12 f) 

shows the detailed construction of each probe. * ' 






ta i 



Shaft speed is measured using two separate probes. A Bently-Nevada Model 300 !| | 
eddy-current probe is mounted thorugh a radial hole to observe a notch 

machined into the outer diameter of the thrust collar. The second probe * , 

consists of a Spectral Dynamics Model SD43-GPT-1 fiber optic probe mounted "^ 

through the turbine casing to observe a single bright mark on the turbine nose fl | 

cone. Both speed measurements are independently connected to readouts and *«' 

automated shutdowns, m . 

7. 3.5 Vibration 

- i 

Two Endevco piezoelectric accelerometers are stud-mounted to the seal end of 

m 

the test rig. One sensor is fixed in the vertical direction and the other in 
the axial direction. The accelerometers are connected to oscilloscope read- 
outs and automatic shutdown devices. " " 



7-12 



7t3t6 Pressures 

Validjme Model DP-15 variable reluctance pressure transducers are used to 
measure fluid pressures » Meter readouts or strip chart recordings are 
provided. Bearing supply pressures are additionally connected to low pressure 
automatic shutdowns. 

7.3.7 Flows 

Both the helium seal supply flow and the outboard seal leakage flow are meas- 
ured using calibrated Venturi flow meters. The flow meter pressure drops are 
sensed with variable reluctance pressure transducers operating in a differen- 
tial mode. The inboard seal flow is derived as the difference between the 
supply and the outboard drain flows. 

7.3.8 Temperatures 

Copper constantan thermocouples are employed. Fluid temperatures are meas- 
ured using sheath-type sensors inserted in the flow stream through 
pressure-tight fittings. Seal ring surface temperatures are sensed using 
small diameter embedded thermocouples, 

7.3.9 Data Acquisition Equipmen t 

Figures 7-6 and 7-13 provide an overall view of the control room and a simpli- 
ii<2d schematic of the data acquisitiun equipment, respectively. Detailed 
schematics of the instrumentation system are provided in Appendix B. 

Oscilloscopes are the principal means for monitoring the dynamic data from the 
test rig. They provide x-y displays of the radial capacitance probes to show 
rotor orbital motion and position with respect to the bearing at each end of 
the rotor and vertical and horizontal seal film thickness at both inboard and 
outboard seals. Dual trace swept oscilloscope displays are provided for test 
rig housing vibration. One speed signal is displayed on a panel meter, while 
the othe.- Is recorded on a strip chart recorder. 



7-13 



IJ 

A Honeywell Model 101 28-channel FM magnetic tape recorder is used to record |J 

all dynamic teat data. A time code is also recorded to permit synchronizing 

the tape recorder with the data logger. I i 



Static data consisting of pressures, tGmpsr.'.cures, and flows are individually 
displayed in the control room using panel meters and a strip chart recorder. 
Some of the panel meters are equipped with alarms to provide automatic 'hut- 
down in selected parameters. 



IJ 



A Fluke Model 2280 data logger is used as the principal means of recording 
test data. All test parameters excepting test rig vibration are included. 

Data signals are serially recorded with a full scan cycle of 2 to 3 s. The' j I 

... , *•! 

value of each test parameter along with the time at which the data scan is 

initiated is transmitted through an RS-232 digital output to a Columbia Data 

Products Model 300D digital cassette recorder tape storage system. Digital 

cassette recordings of test runs are played back and entered into MTl's IBM -r . 

4341 mainframe computer system for analysis. ^j 

ii '^ 











1 



7-14 






ii 
h 



I 

D 



D 
3 



3 
3 




DWGINAL PAGE IS 
B iPDR QUAUTY 




Pi8. J'i Overall View of Heii, 



"n Seal Test Facilicy 



7-15 



1 




Turbtn* Trip 
V«lv« (14) 



l-Nj Manifold 
Bypass Vaiv* 

'2) 



f «• m 

l-Nj Thfuii 

Baaring Supply 

Valva (5( 



PfW»8Ura 

Transductfs 



r- r 




'ig- 7-2 Close-Up Vi 



'" "f "3 M.„ifol. .„, 3„„,, suppl, •;,:.„, 



ORIGINA\. PAGL r 
OF POOii QUAJJTX 



7-16 



c 



D 



D 

U 

D 

U 

U 





D 



D 





D 

U 



Outboaro Seal 

Dfam Control 

vaivs liji 



Main LN] 

Oram valv^ 

ilOi 



n 1 




Fig. 7-3 Close-Up View of 'lesC Rig Drain Valves 



ORIGINAL PA.ir IS 
OF. POOR QUALITY 



7-17 




Fig. 7-^ Close-up View of Helium Seal Tesc Rig 



•1' 



« *l l«. 



ORIGINAL PAGE IS 
Qfi l*OOR QUAUTY 



7-L8 



D 
D 

D 
D 
D 
Q 
D 

U 
Q 

D 

D 
D 

D 


n 



ORIGINAL PAGE T«» 
PCX)R QUALITY 




4 'I 



Fig. 7-5 Overhead View of Helium Seal Test Rig 



7-19 



Dynamic Scoo* 

^OOltO't 



r 



Instrur •nlttlon r't^ ' 

Ampli»(«ft — -^ fL— 

( »*. J B|r Monitor* 




Fig. 7-6 Overall View of Control Room 



OWGnVAl PAGE B 



7-20 



ORIGINAL PAGE !f 
■i POOR QUALITY 




Fig. 7-7 Cl08«-Up Vitw of Sysctffl Control P*n«L 



7-21 



ve-n 



Labyrinth Seal Vent 
(to Atmospher*) 




10 fxm Nom 
Filter 



Helium Seat 

Supply Valve 

(40) 

A 



T0GN2 

Storage 
Trailer 



Run Tank 

Supply Valve 

(38) 



To GN2 
Storage 
Trailer 



GN2 Run Tank 



10 tivn Nom. 
Filter 



FLUID SUPPLY SYSTEM SQ 



/ £UUjgUX £KAM£ 



Fig. 7-8 Fluid Supply System SchemaCi 



i 



^iippiy 
ind6) 



-^ 



Snal Vent 
(to Atmosphart) 
Helium Seal 
Drain Valvo 
(8) 



Tank Vent 
(to Atmosphere) 



T0LN2 
^'^torag^ 

Tank 
Tank Fill 

Valve (17) 

10 fiirt Nom. 
Filter 



® 




(This Line installed 
4th Seal Ttst Only) 



■^H& 



Purge 
(19) 



Manifold 



i40- 



.D um Nom. 
I'ilter 



SYSTEM SCHEMATIC 




Labyrinth Seal 

and Bearing 

Drain Valve 

(11) 



JU 



Tank Level 
Sensor 



Oj-| Pump Bypass 
O^ Valve (16) 



JZg Pump Fill 
**^ Valve (15) 



y \ 



i 





^ tSJU^.UX iLiiAMJi 



PRBCEDIMG PAGF, BLANK HCYT ftLUED 



[ 



upply System Schematic 



7-23 



851621 



I 



TMtMg 

Vibration 



S«al Surtac* 
Temperatures 




PoKlion 



Fig. 7-9 Test Rig lastruaentatioa 



«tl«<4 



Coaxiai 
Connector 



Fringe 

Guard 

(304SS) 



I 




T^i:^ 




Epoxy 

(Eccobond 
104) 



7T-?-r 



1 1 » I K 1^ 



:Ui. 



'.••••.•:v.*:..->;- •-• 
.•-••:.*-.'-*-rf.V-- 




Outer Center 

Sleeve Electrode 

(304SS) (304SS) 






^ V V * V \ V I ■ < ' <" '' ' 



*dii 






»» V. * > ■' ' 



(wJ r^^^ 



V V ' V V \ V 







Fluoroelastomer 

Sealant 

(BEFSET) 



Fig. 7-10 Capacitance Probe Construction 



UD HJ U3 LJ U) l_ 



L_J C-J L 



I 




12 O'clock T-End 

Sea! Probe (ZE11) 

(Size 197) 

12 O'clock S-End 

Sea! Probe (ZE9) 

(Size 197) 



12 O'clock Shaft 

Probe (ZElO) 

(Size 197) 



Axial Shaft 

Probe (ZEIS) 

(Size 125) 




Fig. 7-il 50-9a Heliua Seal Probe Coof Ifuration 



RbNOt fiuARD 



fCNTLR 



TQOQC 




TTflcrSm 



■E 



:v 






Ut 



»6Tl O-tti 

J-i-i- 



CKNTCK oMJCnaaL 






)^^^ 



100CC51 



A /SSUED 



^'?S!!^ 





gMna 


■^M 


MATtRlAi. 


Ol 


IM 


Its 


A«SI 3M MmUK S«U. 


Gt 


OM 


*ts 


MTaHELtM 


03 


OM 


AM 


MCMKL^OA 


6« 


- 


— 


mitit^i. 6da 



MlCROCXJT 

CONNeCTOR 

•OSSOOOI-OOOI 



T 



ia«r M-za 2A RH 

197 PPOBE -(G1)(G2] 



aroM 



ioMfirioi-a 

RCF wax TMO& 



M- 



? 






UNFM-zaZA'RH 

125 PPOBE 



.KSMA 



A. 



Y 



- J 

— CI)> 






IXTAICO 




HOTtS: 

1 2<areoLr«>»i& CuaO - lAa.lnX. AiS 

CTcHfTKATusc KAMc^ - *«aA>F 'tt -ssa*r 

3 UNLCSS ^PC£IFlCALi.y tMEPTED All. 

MATteiAL MUST Bt OK^jOCN C/)MPAneUL 

A rerffiwr BAMbc. - a io eaa i>&a6 

s IOC etFSCT - KL :^7aa c^^ATiNa £<^luti6n 

AMO/a« BlJSSO KjTT»j6 C3MP0ONO TO 

ubkLiic uvu OS uMTt •tx cencMT 

«HD itJtt. exR^tD it* •■ . tti^SET 

ft.^MB COAIIM6 CCHi-ciu-'. 

fc 04 HTMbC OUAMO ExftkSKO HSH nP TO 
nMKAOCX) SCCnbM 




Jk.*. UltX 






26741 



lOOCOSl 



;.sriji jTT H rr 



Fig. 7-12 Capacitance Probes, LOX Seal, and Shaft Probes Configurations 



ri-iiin 



C3 ra C3 C3 d C=! 



PrtampMtkd 
8*nsor Inpuu 



Tachcmetar 
Newport Model 234 



Strip Chan 

Recorder 

Sollec Model 

3318 



S«al OaU 

(SE1) (Pti2 12 18^) 

(TEI2. 13) 



I 




High Seal 

Temperaiurs 

(TF12 13) 



Columbia 
ModelSOOO 



Oala Logger 



Fluke Model 2260 



(SE1) (PE1 -5.10-14.18.20) 
(ZE1 15^ 24) (TE1 -5.10.12 22) 



'Alarm and trip tunclions integral witl) 
meter except lor vibration wtiere lurK:tion 
IS in amplitter. Valtdyrte Model OA-AS) 



Fig. 7-13 SiPH'lifleJ Scheaatlc oC Data Acquisition Kquip«unt 









TABLE 


7-1 






G 








^iIIUMENTATION MATRIX 




Ul 


Sanior 


:odM 




;j9«J<m Haaauramaiff 


fi 
Trip ',r^k SfM'W 


D 


rd«nc . Location T«iC 




m 


< 


i 


(l)Jnl.lri., S-Ead U4. Sh«(c DUpl^ca. 




n 








MorU. 






1 1 


m 






Jnl-lrg.. i End 
Vart. 






LJ 


» 






Jnl.lri.. T-£nd 
Horli. 






•n 


m 


< > 


Jai.Srg., T-End 1 
Vatt. ▼ 







m 


I 


-J 


(2)0eb. Saal tlag Hallua Saal rtia Thick. 


Stffaraat 


al Mathod 


^■v 








U e'alMk 










w 






iMl Imhmi. U a'e. 








n 


mx 






Ub. Saal lla| 








'Q 








12 o'clock 








nu 






Otb. Saal Rial, 3 »'«. 










mi 






Saal Runnar, ] o'e. 








n 


m* 




> 


Ub. Saal Rin(, ) a'e. T 


' • 





au 


i 


i 


Saal Cavlcv.AsUl Dlr. AsUl Shaft Dlap'.aea. 




kj 


nw 




> 


Otb. Saal Rln|, BalliMl Saal TVm Thick. 


Otract Mathod 










10 o'clock 




(aabaddad probaa) 


n 


nu 






Oeb. Saal Rla|, 
12 o'clock 








u 


on 






Otb. Saal Unf, 4 s'c. 










tm 


' ( 


» 


Otb. Saal Rlaa, * o'e. 

1 


1 


" 


n 


m 


r 


i 


TVirblna Whaal Sha^c Spaad 
Thruat Runaar ± 


S,042 rad/«(77K rpa) 
High Alan 


D 


m 










m 






Saal Houalag, Axial Olr. Vibration 


1.0 G paak, 


n 


m 


1 




Saal Nuualas, Vart. Sir. 

Ul2 Supply ".aolfold Praaaura 


IU|h Alam 


D 


m 


t 






m 






Otb. Hal'ua Saal Drain 






n 


m 






Nala In. Orala 






u 


It* 






lr|./Uiiy. Saal Drato 






kJ 


pa 






Hallua Supply Manifold 






■^*% 


mo 






Turblna tnlac 






n 


nu 






Laby. Sa.tl Supply 






u 


nu 






Hallua Saal Supply 








nu 






Jal.lrg. lupply 




).4S MPa (SOO pal() 
Low Alara 


n 


m* 






Thruat lr:i. Supply 




1 


u 


mi 






Otb. Hal. Saal Drain Flow 






mo 






Hal. Saa.'. Supply Flow 


f 




n 


m 






Ul2 Supply Manifold Taaparatura 




u 


TB 






Otb. Ha.'.tua Saal Drain 








m 






Main Irf. Drain 






n 


TW 






Br(./U>y. Saal Drain 






IJ 


m 


> 


< 


Hallua Supply ^unlfold 






^mmS 


TM 


I 


1 


Jal.lrc. Racaaa, S-lad 






*— 1 


TIT 






Jnl.Brf. Racaaa, T-Cnd 






n 


TU 






Thruic Brg. Racaia, 
S-End 






u 


m 


' ' 


> 


Thruat Sr|. Racaaa, 

T-End 






ri 


TOO 


E i 


i 


Turblna talat 






II 


mi 




laby. Saal Supply 






u 


nu 


r 




Otb. Hallua Saal Rin| 




4*C (40*n 




mi 1 


^ 


> 


lab. Hallua Saal Rln( , 


1 


High Alarm 


D 


11) S-tnd r«l 


•r* CO s« 


al 


End ^l Taac Rl« . 


ORIGINAL PAGE IS 


IJ 


T-Fod r* 
(2) Clock BOi 


'•ra CO Tu 
Uiona «i 


rblna End of Taac Rl(. 

a rafarancaa co j tn 


OF POOR QUAUTM 


^aam 


vtawln* 1 


rate Rl( 1 


ro« Saal End. '"■ 


JU 






D 



TABLE 7-2 



MTI ACCUMEASURE^'^ SYSTEM 1000 
REQUIREMENTS AND SPECIFICATIONS 



Requirements 

Probes: (22) All Special Design 

Anplifiars: (II) AS1023-PA Probe Amplifiers 

(1) AS1041-SA Summing Amplifiers 

(2) AS1032-MD Analog Display Unit 

Housings: (2) ASlOll-H InsCrument Housings 

Specifications 

Probe Amplifier (AS1023-PA) 

Linearity: ±0.3% of Range, 10-lOOX Range 

(25-ft Cable) 

Frequency Response: 3 db at 3 kiis 

Output Noise: 40 mv Pcaic-to-Peak at Full Seal* 

Probe Voltage: 5 V rms Maximum 

Output Signal: tlO V dc, lOO-ohm OuCpuC R«flisCanc« 
Summing Amplifier (AS1050-SA) 

No. of Channels: Two Channels Summing to One Output 

Cain: Unit tO.lZ 

Output Signal: ±10 V dc 
Analog Display Unit (AS1032-MD) 

Meter: 0-lOOZ Vertical Scal« 

Accuracy: 2Z 



7-31 



8. C TEST PLAN FOR THE RAYLEICH-STEP, HELIUM BUFFER SEAL 

The program calLed for the testing of the foLLowing two types and sizes of 
seals: 

1. 50-mni and 20-mm spiraL-groove LOX face seals 

2. 50-min and 20-mm Rayleigh-step, helium buffer floating ring seals. 

Due CO budgetary constraints, only the 50-mm helium seals were tested. In 
accordance with the final test plan, four seal packages ';ere installed and 
tested. 

8.1 Test Description 

Four types of tests were to be conducted on the seal packages consi5ting of: 

1. Normal steady-state demonstration tett runt 

2. Acceleration test runs 

3. Test runs with axial runout built into the LOX seal mating ring 

4. Test runs with axial motion imposed on the shaft 

ALL runs were to be conducted in LN2 and repeated in LOX after verifying the 
operational integrity of the test rig and seal package. Because the testing 
was cut back to include only the helium seals, several of the above program 
features were eliminated. The test runs imposing axial runout and externally 
applied axial motion were eliminated because the motion was not relevant '.0 
the operation of the helium seals which are sensitive only to shaft motion in 
the radial direction. Secondly, because the seal working fluid was helium, 
the use of LOX to energize the test rig bearings was not necessary. Thus, LN2 
was used instead, eliminating the need to repeat runs and greatly reducing the 
risk of fire. 

8.1.1 Steady-State Tests 

These tests were planned to prove the basic operational ability of the seals. 
This provided the opportunity to test the seal's endurance and to exper- 



8-1 



ifflcntally examine Che parametric relationships governing actual performanct* 
The former addressed Che specific program objective which requires one hour of 
cumulative test time on each of three seals with at least one half of the total 
test time at a shafc speed of 7329 rad/s (70,000 r/min) and a supply pressure 
of 1379 kPa absolute (200 psia). The latter provided the necessary basis to 
verify the theoretical analyses used in the design stage, also a requirement 
of the program. 

While the design requirements of 1379 k.Pa absolute (200 psia) supply pressure 
and 7329 rad/s (70,000 r/min) shaft speed were clearly the operational goals 
of the testing, preliminary operation at lower speeds and pressures was neces- 
sary Co gain a clear undersCanding of the mechanisms at work and to minimize 
the risk of damage either to the seals or the test rig. Thus, the testing was 
conducted in a step-by-step manner proceeding from relatively safe operating 
points to conditions more and more demanding of the seals. Figure 8-1* shows 
the seal operating map which illustrates a typical sequence of test points. 

The principal test variables were helium supply pressure, speed, inboard drain 
pressure, and outboard drain pressure. Their selection is discussed below: 



^Figures are presented consecutively, beginning on page 8-14. 



8-2 



I 

u 

D 

I 

I 
Q 
B 

D 




1. Helium Supply Pressure. A full range of supply pressures was planned for |f 
each speed going from a minimum of slightly over the inboard drain pres- 
sure, about 317 kPa absolute (73 psia) to the full design pressure of 1379 
kPa absolute (200 psia) The progression of supply pressures generaly 
went from low to high at given operating speeds. As the test speeds 
increased, so did both the minimum and maximum value of the supply pres- 
sures . 



D 




{] 



2> Speed. Because low speeds do not favor the development of good hydrody- 

namic films, the minimum dwell speed was arbitrarily set at 3663 rad/s T | 
(33,000 r/min). During all starts the tester was to be quickly brought up 
to this speed before steady-state operation was attempted. The speed was 



D 
D 

D 
D 



th«n CO be incre«std in increments of S24 rad/s (SOOO r/min), taking data 
at each dwell point up to the design speed of 7329 rad/s (70,000 r/min). 

In practice, dynamic probL^ms in the tester restricted safe operation to 
between 4712 rad/s (43,000 r/min) and S23S rad/s (50,000 r/min). Thus, 
for Che first three seal sets 3665 rad/s (35,000 r/min), 4118 rad/s 
(40,000 r/min) and 4712 rad/s (45,000 r/min) were selected as the primary 
dwell spi:<>^ds. Because of test rig modifications, it was hoped that Che 
Courch seal cast would achieve higher speeds. Actual operation did extend 
Co 5968 rad/s (57,000 r/min); however, higher speeds were precluded by a 
seal failure. 

3. Inboard Drain Pressure. Since Che helium flowing through the inboard seal 
mixed with the adjacent LN2 bearing leakage flow, its pressure was the 
same as the main bearing drain of Che test rig and was thus dependent on 
Cei^C rig operacion. Actual values were 317 kPa absolute (75 piia) 
although variations of ±35 kPa (5 psia) were observed. After the addition 
of the labyrinth seal prior to the fourth seal test, the drain pressure 
vas boosted to 637 kPa absolute (92 psia) because of the need to maintain 
an intermediate pressure in the newly created cavicy. 

Lacking arbitrary control of the inboard drain pressure posed several 
problems. The principal concern was that it limited the minimum supply 
pressure that could be applied. From Figure 8-1, it is easily seen that 
the range of drain pressures clearly eliminated most of the "safe" region 
for lower test speeds forcing most of the operacion into the "high fric- 
tion" region. Secondly, while the frictional forces were a function of 
the absolute supply pressure, the leakage flows were more closely associ- 
ated with Che pressure drop across the seal. The high drain pressure thus 
presented a built-in imbalance resulting in a high-friction, low-flow 
condition at any supply pressure. 

4. Outboard Drain Pressure. The helium exiting the outboard seal went to a 
separate, independently controlled drain. The pressure drop across this 
seal could thus be controlled to the same value as the inboard seal or 
increased up to Che gage pressure of the supply. The latter would tend to 



8-3 



8.1.2 AcceleraCion Tests 

These tests were planned to demonstrate the ability of the seal", to survive 
repeated start-ups at high rates of acceleration with full supply pressure 



applied at the start of the run. The program required two seal set? to be 

subjec 

ft/s^) 



The helium supply pressures applied during the fast starts were increased over 
the course of the two test series. The first series conducted on Seal Set No. 
2 were at a pressure of 931 kPa absolute (135 psia). Those on the third ^eal 
set were done at several pressures progressing frf)m 1069 to 1482 kPa absolute 
(155 to 215 psia) . 

The planned operating scheme consisted of switching on the solenoid operated 
trip valve (14), accelerating the test rig and then switchine. the valve off 
using an overspeed circuit. This resulted in a controlled acceleration 
followed immediately by a coastdown. 

Because the acceleration runs would be very brief, no steady-state data could 
be taken to assess the condition of the seals to determine if any damage had 
taken place. Thus, periodic conventional starts were planned to allow the 
tester to run for a period of ^5 min and achieve steady state conditions. 



8-4 



i 



increase th^2 flow and morn closely approximate the operating conditions I 
envisioned during the design. Thu;i, it was planned to increment the 
outboard drain pressure from 517 kPa absolute (/5 psia) to 103 kPa abso- 
lute (15 psia) during the testing. This permitted the study of a mucb 
wider range of outboard seal flow conditions than could be achieved with 
the inboard seal . 



t 

h 



2 ! I 

subjected to at least 50 starts each at an average rate of 152 m/s (500 *- ' 



!.] 



Because of the tester dynamics problem, the acceleration runs which were 
conducted on the second and third seal sets were terminated between 4188 and I I 
4712 rad/s (40,000 and 45,000 r/min.) A- . starts were at a nominal rate of 500 
ft/s resulting in an acceleration time of about 0.7 to 0.8 s. The acceler- 
ation tests were conducted after the completion of the steady-state runs. 



Lj 




A -- i 



1 • 



IJ 



8.2 Tesi: Schedule s 

Tables 8-1 to 8-6''' present the test schedules describing the steady-state runs 
for Seal Sets No. 1 - No. 4 and the acceleration tests tor Seal Sets No. 2 and 
No. 3. 

The test schedules were based on several important considerations: (1) the 
final test plan (2) the results of the test rig checkout tests and C3) the 
results of the ongoing seal tescs. The final schedule for each seal set was 
written just before the test and tailored to the latest test results. This 
was done to maximize the experimental yield and minimize the risk of damage to 
both the seals and the test rig. Some changes were also made during the actual 
testing. These consisted primarily of eliminating data points at ^.onditions 
of high vibration although several other insitu modifications were also imple- 
mented. 

8.3 Test Procedures 

3.3 .1 Pre paration of Test Kig atJlTI 

All test rig parts were checked for dimensional and material conformi''y in 
accordance with MTI ' s quality assurance program. All test seal parts were 
given a complete dimensional inspection by Stein Seal Company at their facili- 
ties in Philadelphia, PA. The inspection reports are fully documented and 
available for NASA review. 

The test rig was completely assembled including the instrumentation and the 
50-mm helium seals to make sure there were no problems with anticipated 
fitting of parts. The rig was then disassembled, cleaned, reassembled, crat- 
ed, and shipped to the test lab. 



'"'Tables are presented consecutively, beginning on page 8-17 



3-5 



t 

8.3.2 Pr«pT«tion of Tt«t Facility and Ttit Rig at Wyla Laboratoritt | ) 

Q 






A cast sice was prepared including all cesc loup componencs« Cransduccrs, 

concrol room, and conneccions Co Che required fluid storage facilities. An 

area for the test rig assembly, inspection, and parts storage was also set up. 

All test Loop instrumentation was calibrated in accordance with the procedure 

in Section 8.3. S prior to installation in the test loop and at various times 

during the course of the testing. All test loop components were thoroughly 

cleaned. The tester was then assembled in the test loop. All equipment was 

thoroughly checked out and debugged, including verifying all instrument f r 

connections to the concrol room and exercising all controls. 



6»3.3 Teat Facility Operating Procedures 



8-6 




D 
D 



During the design and checkout of the test facility, a detailed operating 
procedure was developed. This covered all steps from initial energizing of 
the facility to the point of establishing stable Minimum speed operation, and 
from Che point the test rig was shut down to the final shutdown of che facili- 
ty. The procedure during the runs was covered by the individual cast sched- If 
ules. The operating procedure follows: 



B 





1. Open LN2 supply line (# 62) 

2. Pressurize main LN2 storage Cank Co 0.17 MPa (25 psig) (# 61) 

3. Open LN2 run tank fill valve* (# 17, * 45) 
U. Verify cank filling. 

5. Record CN2 crailer pressure 

6. Record CHe trailer pressure 

7. Record CN2 shop supply pressure 1 | 

8. Record vapor bulb pressure 0.21 MPa (30 psig) 

9. Lock off vapor bulbs. 

10. Record flow meter sizes 

11. Open CN2 shop supply valve Co manifold (#58) 





D 



*A,B indicates valve to be operated. Valve A is principal valve; Valve B, li 
where applicable, conCrols accion of Valve A. 






12. Verity control panel switches in pre-i eyt po-jition 

13. Activate control panel using key switch 

14. Verify main alarm indicator light is in trip mode 

15. Set all motorized ragulators and valves with momentary switches to 
pre-test conditions 

16. Verify transducers zeroed and B-nuts torqued 

17. Verify transducer isolation valves open 

18. Turn on and verify operation of all instruments at test site 
I'J. Verify all alarms are in latching mode 

20. Open and verify purges: labyrinth seal •:upply, LN2 manifold, test seal 
cavity, turbine. 

21. Open pump fill solenoid valve (#15, #44) 

22. Close pjmp bypass pressure control valve (#16) 

23. Close I,N2 supply manifold bypass pressure control valve (#2, #23) 

24. Allow 15 min. for purge of LN2 manifold and tester 

25. Verify deluge hand valve (#52) closed 

26. Verify GN2 turbine run tank is at psig 

27. Perform ambient accelerometeT tap check 

28. Verify manual turbine speed control regulator (#31) is at zero pressure 

29. Verify turbine control valve (#12) and turbine trip valve (#14) are closed 

30. Verify power plug disconnected from turbine trip solenoid (#29) 

31. Verify GN2 run tank regulator (#39, #34) closed 

32. Verify GN2 supply trailers (2 each) are open 

33. Slowly pressurize GN2 turoine run tank to 100 psig using hand regulator 
(#34) on dome loader (#39) 

34. Verify no flow to turbine 

35. Verify GHe supply pressure regulator #25 is iully closed 

36. Slowly open GHe ullage and verify no flow to P5 or Pll 

37. Open GHe supply from trailers 

38. Close purge labyrinth seal supply and purge test seal cavity. 

39. Open labyrinth seal supply (#41, #24) until Pll is MPa (10 psig). 

40. Turn on hydraulic breaker and pump. 

41. Verify hydraulic supply pressure is 0.52 MPa (75 psig) 

42. Open GHe supply regulator (#40, #25) until P5 is 0.28 MPa (40 psig) 

43. Start Soltec recorder 

44. Adjust GHe manual control (Z) pot until P12 is 0.21 MPa (30 psig) 



8-7 



U' 



I 






Q 
D 



45. Verify P12 ■ 0.21 HP* (30 psig) I, 

46. Verify LMa run cank full 
A7. Clot« purge lolenoid (3) LII2 OMnifold 

48. Verify tank level control tetting 

49. Fully open pump bypass valve (#16) and verify 

50. Fully open LII2 supply manifold bypass valve (#2, #23) 

51. Close valves #S and #6 

52. Clear test pad area 

53. Turn on bootc puap breaker 

54. Start boose pump fl I 

55. Increase LM2 supply manifold pressure PI to 0.28 HPa (40 psig) using pump 
bypass pressure control valve (#16). 

56. Verify no fluid leaks 

57. When Tl is less than -18S*C (-301*F), begin cool down of tester by opening 
valves #S and #6 

58. Verify tester chiUdown, record T6, T7, T8, and T9. 

59. Perform a cold turbine torque check 

60. Perform a cold accelerometer tap check 

61. Close pump bypass pressure control valve (#16) completely I | 

62. Verify varidrive speed control at minimum 

63. Verify turbine trip solenoid valve switch is closed 

64. Verify turbine trip solenoid valve (^14, #29) is closed 

65. Increase setting of labyrinth seal supply regulator (#24) on dome loader 
(#41) until PU is 0.S2 HPa (73 psig) 

66. Turn on varidrive breaker 

67. Clear test pad area 

68. Start varidrive 

69. Fully close bypass valve #2 

70. Increase varidrive speed to set PI to I. HPa (ISO psig) 

71. Increase pressure in CN2 turbine run tank to 2.41 HPa (330 psig) using 
hand regulator (#34) on dome loader (#39) 

72. Verify no flow to turbine 

73. Increase GHe manifold P3 to 1.04 NPa (130 psig) 

74. Open drain valves air supply valve (#18) 

75. Verify controller #68 and #69 at proper pressure I I 

76. Verify that P2 ■ P3 psig 



ti 

n 





D 

D 



8-8 



D 




77. Verify th«C P12 • P2 ♦ 0.21 HP« (30 ptig) 

78. Perform cold curbin* corqu* check. 

79. Plug in power lead co turbine crip solenoid 

80. Clear ceic pad area 

81. Ilowly increase varidrive speed to bring PI to 4.14 HPa (600 psig) 

82. Verify P13 and P14 are at least 3.80 HPa (330 psig) 

83. Verify P2, P3 and P4 are at their proper pressures 

84. Verify P6, P7, P8. P9 are at their proper pressures 

83. Zero capacitance probe amplifiers 

84. Verify Honeywell tape recorder ready 

87. Verify Fluke data logger and Columbia digital recorder ready with nornal 
scan group sequence 

88. Start data logger/digital recorder 

89. Verify all test personnel ready 

90. Verify turoine control mode switch set to center-off position 

91. leset all alarms except turbine trip 

92. Verify alarms are set by observing main alarm indicator light 

93. Set manual speed control valve (#26) to minimum 

94. Verify manual speed control pressure is zero 

93. Verify auto speed control potentiometer set to provide minimum speed 

96. Verify auto speed control pressure is zero 

97. Announce "ready to start" 

98. Start Honeywell recorder 

99. Switch turbine trip solenoid to ON position 

100. Switch turbine speed control mode to manual and slowly increase manual 
control pressure until turbine starts ro rotate 

101. Verify Si and S2 operation 

102. Using manual speed control switch, ramp speed to 3665 rad/s (35,000 
r/min) 

103. Verify manual control pressure are equal 

104. Switch turbine control mode to auto 

103. Using auto speed control potentiometer, adjust turbine speed to 3665 
rad/s (35,000 r/min) 

106. Verify stable operation 

107. Adjist CHe manual control (Z) pot for 0.0 VDC 

108. Increase P3 to 2.41 HPa (330 psig) 



8-9 



i 



I 



11 



r 1 

1.1 



u 

I ^ 

109. Kollow scheduled test plan LJ 

110. Adjust GHe manual control (Z) pot tor 0.5 VDC 

HI. Turn turbine trip solenoid (/iHA, i^25) off I i 

112. Turn turbine control mode switch to center-off position. 

113. Open purjn* solenoid (2) turbine | | 

114. Set manual spped control valve (#26) to minimum 

115. Allow all recorders to run until speed reaches zero 

116. Switch Honeywell recorder and data logger/digital recorder off 

117. Decrease varidrive speed to minimum 

118. Perform cold turbine torque check 

119. Switch varidrive off 

120. Switch boost pump off 

121. Unplug power lead to turbine trip solenoid 

122. Close drain valve air supply solenoid (^^18) 

123. Verify P2 , P3, PA at zero pressure 

124. Decrease setting of labyrinth seal supply regulator (#41, #24) until Pll 
is 0.06 MPa (10 psig) 

125. Ve-ify P12 » 0.21 MPa (30 psig) 

126. Verify supply .anifold bypass valve (#2, '•23'> and pump bypa-i-i vaivt; (#16) 
are closed 

127. Close pump fill valve (#15, #44) and tank fill valve (#17, #45) 

128. Open purge solenoid LN2 manifold and test seal cavity 

129. Adjust GHe manual control (Z) pot until P12 for 0.0 V dc 

130. Switch Soltec recorder off 

131. Shut down hydraulic system 

132. Fully close GN2 run tank regulator (#39, #34) 

133. Fully close labyrinth ^eal supply regulator (#41, #24) 

134. Depressurize GN2 run tank and verify zero pressure 

135. Close GHe supply pressure control regulator (#40, #25) 

136. Close GHe supply trailer valve & GHe ullage 

137. Open purge solenoid GN2 seal 

138. Return control panel to pretest condition (except for purge valves and 
bypass valve (#2 and #16) 

139. Purge overnight 

140. Close all purge solenoids 

141. Turn control panel key switch off - remove key 



8-10 



L-j 



I! 



8.3.4 Pre- and Post-Test Inspection and . nsembly Procedures 

Prior 'bO the testing of each seal set, the test rig was removed from the test 
facility and taken to the assembly area. Here, under clean conditions, the 
seal rings and runners were installed. The assembled seal ring clearances 
were then checked using the capacitance probes by manually moving each ring to 
the extremes of its motion. These were logged and compared with the previous- 
ly measured dimensions. 

After each test series, the test rig was again taken to the assembly area. The 
seal assembly was exposed and taken apart in a step-by-step manner. As each 
piece was taken out, its condition was visually assessed. Also, the seal 
housing was observed for signs of wear debris and contamination as the disas- 
sembly proceeded. When disassembling the seals, after the failures of the 
first and third seal sets, the seal rings were found to be tight on the runner. 
To minimize any additional damage due to disassembly, the runner was chilled 
using LN2. This caused the runner to contract and the seal rings to loosen 
allowing them to be easily removed. On site, photographs were taken of the 
damaged parts. Observations made during testing, assembly and particularly 
disassembly activities were verbally recorded on a tape recorder and later 
transcribed in a written test log. 

After the parts were returned to MTI in Latham, New York, the bore of the 
damaged seal rings and the outside diameters of the runners were measured for 
comparison with the pretest condition. Also, additional photographs were 
taken of the damaged parts including magnified views. 

8.3.5 Instrument Calibration 

Equipment used in the acquisition of data was calibrated, evaluated, main- 
tained and controlled to ensure its accuracy, stability and repeatability in 
accordance with MTl's quality assurance program which is based on MIL-Q-9858A. 
The evaluation results were documented. The evaluation required was dependent 
on the type of equipment and its intended use. 



8-11 



8.3.5.1 Commercial Equipment . Commercial equipment for which sufficient 
information was available relative to its accuracy, stability, and repeat- 
ability were not evaluated if used according to established practices. Howev- 
er, the equipment was calibrated and the results documented. Included in this 
category were all pressure transducers, thermocouples, accelerometers, speed 
pick-ups, instrument preamplifiers, and readout instrumentation. 

8.3.5.2 Special Instrumentation . Specially designed equipment was evalu- 
ated. The equipment was checked out prior to actual use by using actual test 
procedures and conditions to verify the suitability of the equipment for use, 
adequacy, stability, and repeatability. The capacitance probes used for film 
thickness and rotor motion measurement fell into this category because of 
their special size, material and use in cryogenic fluids. Capacitance probe 
calibration was ac unplished by generating a probe gap versus output voltage 
curve. This established the sensitivity, range and linearity. Gap versus 
voltage curves were generated for all probes at room temperature in air. 
Additionally, those probes which were to operate in liquid nitrogen were cali- 
brated in that fluid at atmospheric pressure. Cold calibrations simulated 
actual operating temperatures and fluid dielectric constants. The ability of 
the probe to withstand the large temperature transients was tested by repeated 
immersion in liquid nitrogen after which a visual inspection and both cold and 
ambient gap versus voltage curves was m^de. 

Calibration pi.'ocedures, records, and evaluation documentation on data acqui- 
sition equipment were maintained. This applies to instruments provided inter- 
nally by MTI and to instruments provided by the test lab. This information is 
available to NASA upon request. 

8.3.6 Data Reduction 

During the testing, data were recorded In three ways: digital tape cassettes, 
analog magnetic tape, and strip chart recordings. 

A Columbia Model 300D digital tape storage system connected directly to a 
Fluke Model 2280 data logger serially recorded all test data, transducer iden- 
tification codes, and the time of each data scan. This was the principal 



8-12 



means of recording data throughout the testing. The data logger was triggered 
to start recording before each test run and contii • '.'<1 mtiL after the testing 
stopped. The approximateLy 40 transducer signals were completely scanned and 
recorded every 3 s. Every fifth data scan was additionally printed out on a 
paper tape to permit preliminary data analyses and provide a back up for the 
digital cassettes. Figure 8-2 shows a typical scan. 

The tape cassettes were returned to MTI after the test series for computerized 
reduction. The tape cassettes were played back through an identical tape 
storage system and into MTl's IBM 4341 mainframe cotrnuter. The data were 
filed and output generated in various tabular and graphical forms for analysis 
and comparison with theoretical predictions. In several cases, theoretical 
relationships were simultaneously plotted with the experimental results. 
Figure 8-3 shows a typical computer generated plot of this format. 

Analog magnetic tapes and strip chart recordings provided additional documen- 
tation of test data. Whereas the data logger sampled each signal every 3 s, 
data were continuously recorded on parallel channels on the analog and strip 
chart recorders. This was necessary to capture key data which varied rapidly 
during transients such as the acceleration runs. Strip chart recordings were 
taken of speed, seal supply pressure, drain pressures, flows, and seal temper- 
atures during each run. 

Magnetic tape recordings of speed, capacitance probe output, rotor and rig 
vibration were also made. The tape recordings were reviewed and analyzed to 
evaluate the dynamic behavior of the seal rings. Oscilloscope photographs 
were taken for a permanent record. 



8-13 



SEAL OPERATING MAP 



00 

I 



(A 
£1 

10 






(X 
in 

UJ 

or 
a 



a 
a 

X 
3 



Ui 

I 




iOOC 



2000 



3000 4000 5000 
SHAFT SPEED (rad/s) 



6000 



7000 



8000 



Fig. 8-1 Typical Seal Operating Sequence 



ncfinnacaaaorraa 



OJUGINAJ. PAGE IS 
EPOR QUALITY 






iN 'a- ' 



' -N a- 






L 

c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 
c 



I, 

c 

c 
c 

c 

r 



1 SshL C'ShIn ■ Ti 

2 BRG :Rri:N vTJ 
~ LhS BRG DPN' T4 
4 HE- rinN ■ !■:■ 

; TUR5 5uP '.TlO 

9 3 SE^L SUP' 'il, 

7 T 5Er^t SUR' *i7 

3 S SHT HSG ' T14 
? T SHT H5G' T;l? 

10 S MED HSG':Ti9 

11 T MED H5G'.''17 

12 S LNG HSG'.TIS 
1" "T LNG HSG' Tl? 

14 S SEhl DRn T2.? 

15 T SEhL DRN'-21 
IS HE- SUPCpiU' T22 

20 LN2 MrtN '.Plv 

21 iEr^L DPhIN ■ PL 

22 BFG DRmIn =" 
2" LnB'V dRGDRN' P4 
24 HE;. MrtN • P' ■ 
2'? TiJRB 5vP ' Plti ■ 

30 LtiEY SUP • =11 

31 he:. jEhi. ' =12 

32 ; BRG 5ij= =17 

3" T BRG .-:..P ■ P14 

3" I DRN P'^'jU' Fii 

3? S 2<Jr PLOU' =20 
sirrr, c • 



41 

42 

43 

■if 

4~ 
4>' 
4:- n IE-;. Pn 

50 B P - 1 nG 

51 h:^s 7Ht-=- 



I I.I ;■--? 

1 n - n«-i' 



■' J E — — ' ' 

T I'^TnG 

O ■ Z ' KJl" 



.4 
■ ' "i 



52 P ^^i.'EP 



- iS. -.' iL ^ 



270. 


^ 


DEG F 


23. 


1 


DEG F 


2"4. 


4 


DEG F 


300. 


^ 


DEG F 


7?. 


5 


DEG ' 


4?. 





DE3 = 


4?. 


<; 


DEG F 


3?. 


J 


DEI' F 


10. 


A. 


DEG F 


3o. 


6 


DEG F 


w • 


m, 


DEG F 


14. 


4 


DEG F 


.« .^ 


"• 


DEG F 


.* ^ 


'-I 


DEG F 




5 


DEG F 


5. 


■: 


DEG F 


2?. 


Cl 


DEG F 


612. 


< 


PSIG 


V V 


2 


PII G 


7 1^ * 


e- 


= JI G 


45. 


4 


pj :g 


1'4, 


4 


=s:g 


24. 


5 


=:-IG 


1 ' . 


4 


P5IG 


107, 


2 


P S I G 


- »'i K _ 


4 


=.-:g 


575. 


,;; 


=SIG 


1.12 


5 


= SID 


1. 0-; 


2 


P - I !■' 


i. .■*!?? 


- 


S3M 
JI.:E' 


0. J 


~ 


1I_J 


0. 


,; 


■vj • _ J 


0. 


•^ 


rtl ' T 


0. i 


' 


iiIj 


0. 2 


T 


"ii-j 


0. : 





'*' I _ J 


0. - 


i 


M ' r 


0. 4 


f 


'1 * _ J 


0. " 


e 


:••' = =• 


Z". " 


^ 


-vi-f r 



-^'mN^.-E 



=• ^ 5 r ,■ 



'. 05 ^:_j 

t. 23 -Il-J 



S N t' i . n f'l .' r .' 1.. ,-0 2 " 

j'QP=ED jInG_E J'-n 2" 



NC'..: r4 IjlTi: 
NO'-' r- IG:?!: 



Fig. 8-2 Typical Data Logger Scan from Seal Test No. 4 



8-15 



a» 
I 



007 



.006 - 



.005 - 



«» .004 - 






.003 



.002 



.001 - 



517.0 

37. • 

1.965 

50.0 

0.0 

1.0 

2.07 E 

1.66 

4.003 

ZERO 



DOWNSTREAM PRESSURE (NP«) 
UPSTREAM TEMPERATURE (deg C 
SEAL LENGTH (m) 
SEAL DIAMETER (m) 
ECCENTRICITY RATIO 
DISCHARGE COEFFICIENT 
VISCOSITY (P«-») 
AOIABATIC COEFFICIENT (gsM 
NOLECULAR WEIGHT 
SPEED (rad/s) 
THEORETICAL DATA 
EXPERIMENTAL DATA 
SEAL SET NO. 2 
INBOARD SEAL 




400 500 600 700 800 900 1000 1100 1200 1300 1400 
PRESSURE DROP ACROSS SEAL (KPa) 



Fig. 8-3 Typical Computer-Generated Plot 



V^ _^^ t- ■ J >-■■ ■-« i- — ^ L- — « 



c en t:Z: cr 



a 



Data 
Polnc 



SEAL 


SET 


TABLE 8-1 
NO. 1 STEADY STATE 


TEST SCHEDULE 


Shaft 
Spaad 

rad/s 




Hallua Supply 

Prttaura (P12) 

kPa, aba 


Outboard Drain 
Praaaur* (P2 aom.) 
kPa. aba 



Taat Tlma 
■in. 



1 


4188 
1 


«M 


317 


3 


2 






793 






2 


3 






724 










4 


{ 


1 


6S5 










5 


4188 


>9« 










6 


4712 


8M 










7 






793 










8 






724 










9 






653 










10 


\ 


1 


896 










U 


4712 


963 










12 


5235 


963 










13 






896 










14 






793 










15 






724 










16 






633 










17 


1 




963 










18 


5235 


1034 










19 


5759 


1034 










20 






931 










21 






793 










22 






633 










23 


1 




1034 










24 


5759 


1138 










25 


62C2 


1138 










2ft 






1034 










27 






931 










28 






793 










29 






1138 










30 


' 


1 


1241 










31 


6282 


1344 










32 


61 


306 


1344 


1 




1 


1 



8-17 



I 



TABLE 8-1 (Cont'd) 





Sh«fc 


Hallua Supply 


Ourboard Drain 






Daca 


Spaad 


Praaaura (P12) 


Praaaura (P2 nom.) 


Taat Tlaia 


Point 


rad/a 


U>a. aba 


kPa, aba 


nln. 


33 


6806 


1241 


517 


2 


14 






1138 
















1034 
















931 
















793 












' 


' 


1344 












6806 


1482 












7329- 


1482 
















1241 
















1206 
















1069 
















931 
















793 
















1482 
















1620 






' 




7329 


1482 






30 




6282 


1241 






as needed 




5235 


1034 






as needed 




4188 


896 


1 




as needed 






■0- 


896 


5 


17 


as n 


eeded 



L 

D 


D 

D 

D 
D 
D 



8-18 





D 
U 


u 

D 
D 



i 









TABLE 8-2 












n 


SEAL SET NO. 2 STEADY STATE 


TEST SCHEDULE 




D«ca 
Point 


id/« 


Itollua Supply 

Prassur* (P12) 

kPa, aba 


Outboard Drain 
Praaaura (P2 noa.) 
kPa, aba 


Taat Tlaa 
■In. 




3665 


724 


517 


5 








6S5 






1 








586 
















793 
862 


1 

517 












862 


310 












793 
















724 
















635 

586 


310 












586 


103 












655 
















724 












1 


> 


793 
862 


f 

103 








3665 


802 


511 








&188 


862 
















793 
















724 
















655 


' 














931 


517 












931 


310 












793 


1 












793 


♦ 












724 
o55 
655 


310 

i 

103 












724 
















793 












' 




862 
931 


i 

103 








4188 


931 


517 


' 


1 




4 


?12 


931 











I 



8-19 



TABLE 8-2 (Cont'd) 





SiMft 


Haliua Supply 


Outboard Drain 






Data 


9p««d 


Prataur* (P12) 


Praaaura (PI noa.) 


Taac Tlaa 


Point 


r«4/t 


kPa, aba 


kPa, aba 


aln. 


14 


4712 


1000 


517 


i 


3S 






862 










31 






7»3 










J7 






724 


i 






31 






6SS 


517 






3« 






6SS 


310 






40 






724 










41 






793 










42 






862 


' 


\ 




43 






931 


JIO 


1 

■ 


44 






931 


103 






43 






882 










4« 






793 










47 






724 


' 






48 


' 


1 


655 


103 






49 


4712 


931 


517 






50 


5235 


931 










51 






1000 










52 






862 










53 






793 


' 


' 






54 






724 


517 






55 






724 


310 






56 






793 










57 






862 










58 






931 










59 






1000 


310 






60 






1000 


103 






61 






931 










62 






862 










63 






793 




• 






64 


' 


1 


724 


103 






65 


52 


35 


931 


5 


17 


' 





8-20 



u 

u 

D 
D 
U 
D 
D 
LI 
D 
D 
D 

D 
Q 
D 
U 
Q 



TABLE 8-2 (Cont'd) 





Shaft 


Itallua Supply 


Outboard Drain 






D4t« 


Spaad 


Prataur* (PI2) 


Praaaura (P2 noa.) 


Taat Tlaa 


Point 


rad/t 


kPa, aba 


kPa, aba 


■In. 




J759 


931 


S17 

a 


1 








1000 
















1069 


517 












862 












862 


310 












931 
















1000 
















1069 














862 


310 












862 


103 












931 
















1000 


f 






57 


59 


1069 


103 




L 



8-21 



TABLE 8-3 



SEAL SET NO. 3 STEADY STATE TEST SCHEDULE 



IJ 

IJ 



Polnc 



Shaft 
Sp«ad 
r«d/> 



Halluiii Supply 

Fraiaura (P12) 

IcPa, abi 



Oucboard Drain 
Praiiura (P2 nom.) 
kPa , aba 



Taic Tiaa 
mln. 



1 


3665 

1 


655 


517 


§ 


2 






793 






1 

M 


3 






931 










4 






1069 


1 1 






5 






1206 


5 


[7 






6 






655 


310 






7 






793 










8 






931 










9 






1069 


f 






10 






1206 


310 






11 






655 


103 






12 






793 










13 






931 










14 


\ 


' 


1069 


} 








15 


3665 


1206 


103 






16 


4188 


655 


517 






17 






793 










18 






931 










19 






1069 










20 






1206 










21 






1344 


517 






22 






655 


310 






23 






793 








24 


' 




931 


1 






25 


4188 


1069 


310 






26 






1206 








27 






1344 


T 






28 






655 


103 






29 






793 










30 






931 










31 






1069 










32 




r 


1206 








1 



^ 

}::' 



L! 



g, , 



ij 






ij; 



i ■ 






8-22 



I J L 



TABLE 8-^3 jCont'd) 



D«ca 
Poinc 



3h«fC 


HtlluD Supply 


Outboard Drain 




9p««d 


Fr«*iur« (P12) 


Praaiura (P2 nom.) 


Taac Time 


rad/i 


kPa. abi 


kPa, aba 


nln. 



33 


4188 


1344 


103 


T 


34 


4712 


793 


517 






35 






931 










36 






1069 










37 






1206 










38 






1344 


' 






39 






1482 


517 






40 






793 


310 






41 






931 










42 






1069 










43 






1206 










44 






1344 


M 






45 






1482 


310 






46 






793 


103 






47 






931 










48 






1069 










49 






1206 










50 






1344 




< 






51 


4712 


1482 


103 






52 


5235 


793 


517 






53 






V31 










54 






1069 










55 






1206 










56 






1344 




< 






57 






1482 


517 






58 






793 


310 






59 






931 










60 






1069 










61 






1206 










62 






1344 


1 


' 






63 






1482 


310 






64 


1 


' 


793 


103 


- 


65 


5235 


931 


103 


2 
1 


66 






1069 










67 






1206 


, 








68 


1 


35 


1344 






1 


1 


69 


52 


1482 


1 


03 




1 



8-23 



TAiiLE 8-A 



F!RAL 



A f'TEADY STATE TliHT r^CHEDUIJ' 



Point 

1 

2 

i 

4 

5 

6 

7 

8 

9 
LO 
1.1 
12 
13 
14 
15 
16 

I ** 
J. / 

18 
19 

:o 

21 

23 
2(. 
2S 
26 
27 
28 
29 
30 
31 
32 
33 



Shaft: 
cad/u 

MMMMMHMMMK4 

3605 
418H 
4712 
S23S 
i?59 
6282 
6806 
'329 



H«lfu« SuppJ.y 

Praitaura (P12) 

kPa, aba 



Oucboard Drain 
Praiaura (P2 nom.) 
KPa, abi 



TaiC Tlma 
'Sin. 



7 329 
6282 



827 


5 


.7 


1 


827 










96S 










1103 










1241 






If 


1379 






1 


13V 9 






\H) 


1241 






1 


U03 










965 
827 


1 1 

517 






1379 


310 






1241 










1103 










963 
827 


1 ' 
310 






1379 


i03 






2141 










1103 










965 

827 


1 


13 






1379 


517 






1241 










1103 










96S 
827 


• 

5 


• 
.7 






1379 


310 






1241 










1103 










965 
827 


1 

3] 


f 

LO 


\ 


> 



% 

*:.. 



8-24 



TABLE 8-4 (Cont'd) 



DaC3 
Point 



Shaft 


Htllum Supply 


Outboard Drain 


9p«td 


Prasaure (P12) 


PrcsBura (P2 nom 


r«d/s 


IcFa, aba 


kPa, aba 



Tast Time 
min. 



34 


6282 


1379 


103 

1 


T 


35 






1241 










36 






1103 










37 






965 


1 






38 


' 


' 


827 


103 






39 


5235 
1 


1379 


517 






AO 






1241 










41 






1103 










42 






965 


1 






43 






827 


517 






44 






1379 


310 






45 






1241 










46 






1103 










47 






965 


1 








48 






827 


310 






49 






1379 


103 






50 






1241 










51 






1103 










52 


{ 




965 


1 








33 


5235 


827 


103 






54 


4188 


1379 


517 






55 






1241 










56 






1103 










57 






965 


■ 


' 






58 






827 


5i7 






59 






1379 


310 






60 






1241 










61 






1103 










62 
63 






965 
837 


3 









64 






1379 


103 






65 


1 


1 


1241 


103 


1 




66 


4188 


1103 


103 

1 


2 


67 




965 


I 




68 


1 


' 


827 


1 


r 

03 




r 



8-25 



Run No. 



1-5 

6 
7-11 

12 
13-17 

18 
19-23 

24 
25-29 

30 
31-35 

36 
37-46 

47 
48-57 

58 



TABLE 8-5 



SEAL SET NO. 2 ACt^ELERATION TEST SCHEDULE 



Accalaraclon Rata 
H-High L- Low 



Max. Spaed 
rad/s 



Supply Prtaaura, 
(n2) kPa, aba. 



Drain Pressure Test Time 
(P2) kPa„ abs rain. 



H 

I, 
H 
L 
H 
t 
H 
L 
H 
L 
H 
L 
H 
L 
H 
L 



4188 



931 



517 



u 

u 

tl 



1 

e 


i 


1 

5 
1 


u 


5 

1 
5 


ii 


1 
5 
1 





5 
1 
3 


if 


1 
5 


y 









D 



>■ I 



11 



8-26 



y 



9.0 REFERENCES 

1. Shapiro, W., Walowit, J., Jones, H.F. "Interim ReportJ LOX Turbopump 
Seals Performance Analysis Verification Analysis, Design of Spiral-Groove 
"^sals." MTI Report 83TR5, prepared for NASA/LeRC, September 1982. 

2. Shapiro, W. , Artiles, A., Jones, H.F. "Interim Report: LOX Turbopump 
Seals Performance Analysis Verification Analysis, Design of 
Floating-Ring, Rayleigh-Step, Helium Buffer Seals." MTI Report 83TR6, 
prepared for NASA/LeRC, October 1982. 

3. Shapiro, W., Dunne, J., Hamm, R. "Interim Report; LOX Turbopump Seals 
Performance Analysis Verification Analysis, Design of Seal Test Rig." MTI 
Report 83TR14, prepared for NASA/LeRC, January 1983. 

A. Hamm, R., Shapiro, W. "Final Test Plan: LOX Turbopump Seals Performance 
Analysis Verification." prepared for NASA/LeB.G, November 1982, 

5. Burcham, R. E., and Boynton, J. L. "Small High-Speed, Self-Acting Shaft 
Seals for Liquid Rocket Engines." NASA CR 135167, Rl/RD77~195, July 1977, 
prepared for NASA/LeRC, by Rockwell International, Rocketdyne Div,, 
Contract NAS3-17769. 

6. Artiles, A., Walowit, J., and Shapiro, W. "Analysis of Hybrid, Fluid-Film 
Journal Bearings with Turbulence and Inertia Effects." ASME Publication, 
Advances in Computer-aided Bearing Design, 1982, p. 25. 

7. Artiles, A., Shapiro, W. , Jones, H.F. "Design Analysis of Rayleigh-Step, 
Floating Ring Seals." ASLE Transactions, Vol. 27, October 1984, p. 
321-331. 

8. Shapiro, W., Walowit, J., Jones, H.F. "Analysis of Spiral-Groove Face 
Seals for Liquid Oxygen." ASLE Transactions, Vol. 27, Number 3, July 1984, 
p. 177-188. 



9-1 



9. Mui jderman, E. A. "Spiral-Groove Bearings." PhilLips Technical Library, 
New Yrok: Springer-Verlag Inc., 1966. 

10. Elrod, H. G. and Kg, C. W. "A Theory of Turbulent Films and Its Applica- 
tion to Bearings." ASME Trans., Journal of Lubrication Technology, July 
1967. 



9-2 



APPENDIX A 



FLOW THROUGH HEUUM SEAL INCLUDING 
INERTIA EFFECTS 



A-1 



NOMENCLATURE 



ci 



A * Annular cross section arcs of sital 

C " Radial clearance 

C- "> Discharge coefficient 

1 - Seal length 

p ■ Downstream ambient pressure 

Intermediate pressure Inmed lately downstream of Inlet 

Intermediate pressure at which flow becomes choked 

Inlet pressure to seal region 

Gas constant 

Intermediate absclute temperature 

Absolute temperature of Inlet gas 

Fluid velocity 

Seal eccentricity ratio 
y - Ratio of specific heats of gas 
li ■ Fluid viscosity 
p ■ Fluid density 



R 
T. 



T 
. s 



PRECEDING PAGE BLANK NOT FILMED 



A-3 



There are two regions of flow to be considered over the sealing land : 

(1) An inlet region where the flow is strictly inertial and is treated 
as an orifice. 



(2) A land region where both inertial and viscous effects are considered. 



The unknown is the intermediate pressure between these two regions. 



D 
U 

11 

ii 






In the inlet zone, the flow is described by an orifice equation: 



q - AC, 



1/0 n - i / - Y-1 > 1/2 



i^ 



where p^ - p^/Pg 



ci 



ci 



■y+1 



T 
Y-1 



If p^ < p^^. set p^ - p^^ 



(choked inlet) 



(2) 



ii 



In the film region the following equations apply: 



dp . dv 

dx ^ dac 



12ul 



C (1 + -J e ) 



q " pv A, p - p/RT 



1^ + 
dx 



M 



d 
dx 



12ul 



c^a+leh ^^ 



(3) 

(4) 

(5) 



T ", 



I } 
^■1 



t 



A-4 



Ajjp'^p 



l3.f ± 

A dx 



(Inp) -~ 



12ul 



C^ (1+1 E^) 



1 
A 



(6) 



Y-1 

P -liT" 



.1/y 



(7) 



n -IZI 



1+Y RTi 



f I±l 



Y+l\ 

- p, v - 



i 



i in 

Y 






12ul 



_ 1 
3 2, A 



C*(l+Y e") 



(8) 



Y+1 



2Y ^ 
1+Y 7i 



iPi -Pa ^ 



^ Kl ^a" p/ c2(l+| e^) ^ 



(9) 



p^2 (,2(1+1 ^2)^ 
24vil RT. • 



q-^ ,3-2RT3(- 



(10) 



Y-1 

2Y Pi "^ 

Yfl Ti \^i 



Y+1 Y+1 

- P, 



-^ Hfel 



(11) 



where T. ■ T./T 



da 



Choking will occur at T=r ■ 



Differentiate above equations with respect to p , set dq/dp - 0, solve for 
p and call the value to obtained p . 

a C * 



•-.2 Y 
S£ 'i 

2Y 



Pi 



I±L 



Y 
1+Y 



(12) 



If p < p , set p ■ p (choked flow) 

cl C a C 



For isothermal flow, set Y« 1 in equations (11), (12). 



A-5 



I) 
11 



The flow through th« Inlat rtglon Is tquatcd to the flow In the film ration 

and th. int.nn.di.t. pr«..ur. p^ is solved for numerically. [ 



1.1 

I ] 



I 



53 I 






] 



1, 

I] 



A-6 



I 

I 

D 

D 
3 



APPENDIX B 



INSTRUMENTATION SCHEMATICS 



B-1 



^ ia» tti^ TOT QHwY^ 



^«UVC Ur>CiJM«K* 4'JUjaWn 



ORIGINAL PAGE 
D£ S'OOR QUAU1> 







niT^.H 



1^ ft t»*»^l I 



_i i^ iy ttt'S9ttfklStf\t"4Tme ** — * — <^IHJ,«T* LM MM.TMT OH-v Ly«»i^v 

©^ (7)(7)(f)(|)(^(V)®@ (mj) iii)v^(a)vii)®®®® 



/kjk 



d 



-^1» 



T« gmthdi. noom 



■ ^ i:85^ r ^^^ i|t\' 



'-STBIP 



}|it»i^i4|i3|>jl^^^[|^^t^jfk|!|l|jit|j»|i^^ 



lltllTlli 



ikt 



UtlltM 



ii 



ilM 



M 




Qr^NAMIC SIGNAL Q^ T CK ft^KlgL 
TO PtADSuTj (SetSHEtTZ) 



CCtUTinuCD SKjNAL IXC 



/ fiOUDOUX ERAME 



-/•-t': 



r.' 



^^c 






^ I 




«i ^^aSf > :a . l fe , ii i ih^l , 



•t-Ct 



(!^©(?)®(|)(S)(^(?)®(|)(g)(5)(^ 



^ iOUJOUl£KAMa 



sru 






n^klkni=nkn^fc-^i^ 



i iiMi#i|^M fHii|tiUHHiiiii 




^^S^&f^' 



i^itMii'' 




g^ffgBpwp agg»gi 



iHiHJSu^^'^ 



Bffiitiim 



^':ttte 



•^ "^ "TsrsrstPN? 



i^xi^ PRECEDING PAGE BLANK NOT FTLMH) 



" Ti Tj T« T« T« T7 T« rii no tI, t'h Tli 
y^yirniir n ^kjmal ttrm imal strip 



Fig. B-1 Detailed Instrumentation Schematic 
Sensors and Preamplifiers 

B-3 



■OUXHIT FRAME 



yfNA'^i: lI'itlAL f-ArZH KkNtL 
■ in 1.1 M ra fK «ti IK IIS |M tn lA m (c to it :« x: <* 



l|m|||||||||n[ i |'[ i || l_^ 






ORIGINAL PAGE 
or POOR QUAMT 



n-m iMranncs c 



CauMHA MOO 



C£MDtTlONC& 5K>hAL - 

p« •? p« •• ■• m >it,ji«sun( M m f« )n *<• *■• •<« 




SPEED ADJUST 



^ fOLDOUT CRAia 



TihM, sr*i£ 



ti\).oi-d.,\^s^ 




^yjAUTY 



■KO-'awo PAGE BLANK N</r FtLUO, 



Fig. B-2 Detailed Instrumdntaticn Schematic - 
Indicating and Recording Instruments