PARAMETERS AND DESIGN

Transcription

1 TURBOPUMP PARAMETERS AND DESIGN SSME Turbomachinery Pump Specific Speed For a pump with M stages:!h t = p t 2 " p t1 M#g head per stage!!v! S = 3 4 stage specific speed ( g"h t ) where:! is the rotor (angular) speed!v is the volumetric flow rate!g is the fluid specific weight! S is a key parameter characterizing: the pump type: centrifugal (! S << 1) mixed-flow (! S! 1 ) axial (! S >> 1) the pump performance: efficiency! p (! max for! S " 1) flow coefficient!

2 Pump Speed and Diameter For a single pump stage ( M = 1): stage flow coefficient: V! =! A U =!V "r T b TURBOPUMP PARAMETERS AND DESIGN ( )(#r T ) stage head coefficient: where: b and:! = g"h t U 2 = "p t #$ 2 r T 2 A =!r T b is the discharge blade height is the discharge cross-section!h t =!p t "g is the total head rise U =!r T is the rotor tip speed!! near design conditions Pump shaft power: SSME Turbomachinery P =! V!p t " (to be provided by the turbine)

3 Pump Selection (based on requirements) TURBOPUMP PARAMETERS AND DESIGN Requirements derived from engine system: fluid (hence density! and viscosity µ ) inlet total pressure p t1 inlet total temperature T t1 pump discharge total pressure p t 2 pump flowrate!v operating range! min,! max Head coefficient is limited for pump type: SSME Turbomachinery!p t "# 2 r T 2 $% max & #r T =!p t % max " Hence the required!p t constrains rotor speed! and tip radius r T Given the required pressure rise, the designer: maximizes! within constraints: generally advantageous to maximize speed many limits restrict max speed selects the rotor tip radius r T and therefore the pump size (scales with r T ) if necessary increases the no. of stages M (lower stage head!h t = ( p t 2 " p t1 ) M#g )

4 Cavitation and Suction Performance TURBOPUMP PARAMETERS AND DESIGN Inlet flow pressure strongly affects pump design Cavitation can develop when: p < p V (vapor pressure) Effects of cavitation: performance: loss of head, efficiency, flow rate machine life: structural damage (erosion) flow instabilities and dynamic loads rotordynamic unbalance (instabilities) Suction performance specified by: NPSH = p! p V (Net Positive Suction Head) "g 2g or by the nondimensional parameter: + u2!!v! SS = 3 4 (suction specific speed) ( gnpsh) ( S S when expressed in rpm gpm 0.5 ft ) SSME Turbomachinery

5 TURBOPUMP PARAMETERS AND DESIGN SSME Turbomachinery Turbopump Inducers Low NPSH requires the use of an inducer: inducer normally necessary if:! SS > 20 ( S S > ) (limit should be lower for long pump life) High suction specific speed inducers: permit minimum pump inlet pressure result in lower tank pressure and weight

6 Turbine Performance Parameters Isentropic velocity ratio: U!r = T c 0 2"#p U i 2! i c 0 = TURBOPUMP PARAMETERS AND DESIGN! ("r Ti ) 2 i 2#$p (single stage) (multistage) Degree of reaction: R =!h rotor " 0 impulse stage = #!h stage $ 0.5 reaction stage Design trade-offs for increased efficiency: increase the turbine tip speed U =!r T : speed limitations encountered due to stress and bearing limits increasing diameter increases weight, envelope, cost decrease the spouting velocity c 0 = 2!"p : decreases turbine power, generally unacceptable: could be traded-off with higher flow rate increase the number of stages, decrease the stage value of c 0 : increased complexity, weight, envelope, cost SSME Turbomachinery

7 TURBOPUMP PARAMETERS AND DESIGN SSME Turbomachinery Rotordynamics Trends influencing space T/Ms rotordynamics: higher engine power densities (high p c ): higher turbopump pressures higher rotational speeds: smaller, less rigid turbopumps Dynamic problems encountered: LOX pump explosions unsteady cavitation problems seal reliability bearing loads and durability turbine blade failures turbine wheel vibration balancing, synchronous vibrations critical speeds start transient problems synchronous whirl

8 SSME TURBOMACHINERY Vehicle/Engine Requirements Two turbopumps required for each propellant: engine must accept low inlet pressures: minimize vehicle propellant tank weight dictates low turbopump speed main combustor must operate at high pressure: maximize available energy requires high pump discharge pressures dictates high pump speed Engine must throttle: dictates operation over wide flow range Other design considerations: efficiency and weight dynamic seal life and efficiency rotor axial thrust balance bearing life rotor balancing capability rotor critical speeds and dynamic stability service in high pressure hydrogen or oxygen

9 SSME Propellant Flow Schematic SSME TURBOMACHINERY

10 SSME Powerhead Arrangement SSME TURBOMACHINERY

11 LOW PRESSURE OXIDIZER TURBOPUMP (LPOTP) Design Description Pump: four blade axial flow inducer Turbine: full admission six stage impulse turbine LOX powered, driven by HPOTP Pump/turbine rotor axial forces balance: residual carried by thrust bearings Operates below first rotor critical speed Design for Oxygen Hazards Close clearances for performance, no rubbing: heat generation, rapid oxidation of metal Protection provided by material selection: K-Monel rotor elements & inducer tunnel liner silver wear rings silver plating at turbine lands

12 LOW PRESSURE FUEL TURBOPUMP (LPFTP) SSME Turbomachinery Design Description Pump: four blade axial flow inducer four partial blades at exit Turbine: partial admission 2 stage impulse turbine LH2 powered, driven by HPFTP Pump/turbine rotor axial forces balance: residual carried by thrust bearing pair Operates below first rotor critical speed Utilizes a lift-off seal: prevents hydrogen leakage external to engine pressure actuated by internal pressures

13 LOW PRESSURE TURBOPUMPS Development Problems LPOTP: early thrust bearing ball wear: reduced loads by force balance change reduced labyrinth seals increased turbine load in aft direction LPFTP: low pump performance: reduce no. of diffuser vanes eliminate partial blades reduced energy absorption from turbine: switch to partial admission turbine labyrinth seal degradation: caused by non-symmetrical admission: resultant radial load implemented symmetric admission

14 HIGH PRESSURE OXIDIZER TURBOPUMP (HPOTP) Design Description Pump section: main pump: double entry centrifugal impeller two inducers volute with vaned diffusers discharges to: LPOTP turbine preburner (boost) pump main combustor preburner (boost) pump: single entry centrifugal impeller boosts ~11% of engine flow Two stage reaction turbine Axial thrust balance: preburner impeller balances turbine thrust balance piston controls residual

15 Main Pump Design HIGH PRESSURE OXIDIZER TURBOPUMP (HPOTP) High speed desired for turbine efficiency High suction performance: low Net Positive Suction Head (NPSH): NPSH = p t1! p V "g high suction specific speed:! SS =!!V ( g NPSH) 3 4 Main impeller sets shaft speed! : determined by suction performance capability:!!v "! SS gnpsh ( ) 3 4 = k # constant SSME Turbomachinery Hence, for single/double entry impellers (!V 1 = 2!V 2 ) with equal suction performance:! 1!V 1 =! 2!V 2 "! 2 =! 1!V 1! V 2 =! 1 2 (41% higher speed for double entry impellers)

16 HIGH PRESSURE OXIDIZER TURBOPUMP (HPOTP) SSME Turbomachinery Balance Piston Axial motion of the rotor used to generate: differential leakage of discharge flow on rotor sides restoring axial force on the rotor Main features: effective only during pump operation decreasing effectiveness at reduced pump loads Bearings Spring preloaded, angular contact pairs Propellant cooled Turbine end 57 mm, axially restrained Pump end 45 mm, axially unrestrained Dynamic seal package: controlled gap, intermediate seal with He purge: separates hot gas and oxidizer leakage flows

17 Preliminary Design Selection HPFTP design requirements: LH2 flow: 1.08 m K inlet pressure: 14.8 bar discharge pressure: bar operational range:!" " max = 0.2 Comparative design features: no. of stages: 2 3 (3 stage design selected) rotational speed (rpm): tip diameter (m) (same in both cases) tip speed (m/s): (compatible with Titanium rotor) efficiency (significant efficiency benefits) stage head coefficient: (reasonable head coefficient) specific speed: (compatible with radial design) Additional features to be checked for speed compatibility: turbine design bearing and seal capability suction performance rotordynamics

18 HIGH PRESSURE FUEL TURBOPUMP (HPFTP) SSME Turbomachinery Design Description Three stage centrifugal pump: two high efficiency cross-overs diffuser and volute at 3rd stage Two stage reaction turbine: 700 HP per blade hydrogen embrittlement protection Angular contact 45 mm duplex bearings: thrust bearing for transient loads Pressure actuated lift-off seal Axial thrust balance: pump impeller balances turbine thrust balance piston is single acting Development Problems Turbine end bearing life Rotor subsynchronous whirl Turbine blade life

19 Turbine Design Selection Turbine design requirements: working fluid: LOX-Hydrogen: selected by engine system based on mission studies low pressure ratio, high flow turbine: engine systems selected stage combustion cycle limits pressure ratio across turbine requires flow increase to achieve power required power for driving the pump: P p =!V!p t " p = 57.8 #10 7 MW Turbine power output (total-to-static, perfect gas): P t =! t!mc p T t1 " T 2 Hence P t depends on: ( ) =! t!mc p T t1 1" p 2, * +, # $ % p t1 mass flow rate through the turbine!m the energy content of the fluid c p T t1! T 2 ( ) the turbine efficiency! t & ' ( () "1) ) - /. /

20 Selection of Turbine Inlet Conditions Inlet conditions established through iterations with engine balance Inlet pressure: increased pressure desirable to minimize flow and temperature results in increased pump pressures requires higher power and torque Inlet temperature: higher values desirable for power lower values desirable for life values 1111 K (2000 R) evaluated Parameters selected after iterative engine system studies: inlet pressure p t1 : MPa (5650 psi) pressure ratio p t1 p 2 : 1.56 inlet temperature T t1 : 1106 K (1990 R)

21 Turbine Design Trade-Off and Preliminary Selection Selected pump rotor speed and diameter: rpm and m: turbine diameter range: m Available energy determined from fluid conditions (spouting velocity c 0 ) Comparative design features (3 stage impulse design selected): no. of stages: tip diameter (m): stage isentropic velocity ratio: turbine type: impulse reaction reaction impulse efficiency Constraints: diameter has a significant weight impact added stages adds length/weight/complexity/cost turbine stress limits diameter Final selection optimized within above constraints: 2 stage reaction blading tip diameter: m

22 Potential Bearing Arrangements 1. Bearing outboard, pump inlet in the middle: requires high pressure hot gas seal permits hot gas flow into pump inlet impacts suction performance 2. Turbine overhung, pump inlet outboard: turbine bearing too large (excessive DN value) shaft sized for carrying torque bearing diameter set by shaft diameter 3. Pump 1st stage overhung, turbine bearing outboard: requires special tooling for 1st cross-over stage no significant critical speed gain 4. Bearings outboard, pump inlet outboard: small bearings (45 mm) yield low DN (1.66!10 6 in rpm ) hydrogen bearing state-of-the-art ( DN! 1.66 "10 6 in rpm ) critical speed control favorable pump-turbine sealing arrangement Selected as optimum

23 Seal Selection All dynamic seals are labyrinth seals operate with clearance maximize seal life no rubbing velocities non impact on shaft speed Pump pressure exceeds turbine pressure: no need for mechanical dynamic seal H2 leak into turbine exhaust used in MCC Seals designed for minimum FPL: highest pressure most critical performance No high pressure external flange seals required: piston rings seals used internally casing area vented to inlet pressure

24 Typical Seal Operating Characteristics No. Fluid Δp Clearance Flow bar mm kg/s 1 LH LH LH Hot Gas 48 * Hot Gas Hot Gas LH2 3.4 * LH2 408 * LH LH LH GH2 143 * * Static seals designed for zero clearance

25 Suction Performance Suction performance requirements: engine Net Positive Suction Pressure: NPSP = 36.7 FPL suction specific speed without LPFTP:! SS = 361 (way too high) current technology limit (optimistic):! SS = 162 (! SS = 20 corresponds to S S! ) speed required to get! SS = 162 :! = 1737 rad/s = rpm (too low for HPFTP) Hence a separate low pressure pump is required LPFTP supplies sufficient head for HPFTP: HPFTP minimum NPSH! 1900 FPL maximum! SS = 13.8 (affordable value)! SS = 13.8 is low enough to design HPFTP without an inducer significant advantage in terms of rotor length and critical speeds

26 Critical Speeds Critical speeds evaluated for compatibility: operating speed range: rpm required margin: 20% from critical speeds optimum bearing support stiffness required pump operates between 2nd and 3rd critical Design Speed Selection Selected speed satisfies all requirements: provides high efficiency with 3-stage pump configuration provides high efficiency with 2-stage turbine configuration impeller and turbine tip speeds within structural limits bearing DN value within state-of-the-art limitation for life no high seal rubbing velocities suction performance capability of pump provided critical speed margins met

27 Axial Thrust Balance HPFTP axial thrust requirements: design for axial thrust control minimize axial loads carried by bearings bearings usually required to take start-up load Design options available: arrange turbine thrust to oppose pump thrust locate pump wear rings to minimize pump thrust use self-compensating balance piston optimize sump location

28 Diffuser-Crossover System Critical design feature: must provide efficient diffusion (8340 m/stage): from 9780 m (impeller discharge) to 1350 m (crossover discharge) space/weight must be minimized: requires tight flow turns difficult fabrication Three concepts have been proposed and tested: radial diffuser and crossover vane system axial diffuser and crossover vane system continuous channel diffuser and crossover Tested in air rig using existing impeller: axial diffuser showed stall, low efficiency radial diffuser had highest efficiency (75.3%) continuous channel diffuser (selected): comparable efficiency (74.0%) easier fabrication, less structural problems

29 Fluid Properties Impact on Design Oxidizer generally highly explosive: must avoid rubbing contact results in larger clearances results in lower pump efficiency requires in-depth analyses throughout operation Hydrogen compressibility results in loss of head: LH2 thermodynamic properties more sensitive LH2 results not following normal scaling laws predicted performance dependent on thermodynamic modeling High vapor pressure fluids provide thermodynamic benefits: cavitation is delayed due to local thermodynamic effects referred to as thermodynamic suppression head required NSPH is reduced effect can be very significant in LH2: 3:1 NSPH reduction if pump velocity is sufficiently low

30 Rotordynamic Design HPFTP rotor architecture: single shaft on 2 supporting end bearings 5 rotors (3 pump stages, 2 turbine disks) 12 labyrinth seals Elastic supports: radial flexure for shaft/bearing/carrier Belleville damper of flexure Belleville spring axial preload on bearings Operation: operating speed range: rpm required margin: 20% from critical speeds optimum bearing support stiffness required pump operates between 2nd and 3rd critical

31 Subsynchronous Whirl Problem Prior experience: large steam turbines, blowers, compressors bearing whip textile spindles High incidence in LH2 space T/P: early MK 15F (J-2 Program) MK 9 (Rover Program) MK 25 (Phoebus Program) 350K P&W (USAF) Eliminated by stiffening of shaft and bearings Possible contributors to rotor instability: internal hysteresis turbine (aerodynamic) cross coupling seal forces nonlinear effects axial-radial coupling (balance piston, rotor/casing)

32 Designs for Increased Stiffness and Damping Bearing stiffness (on the right): increases from configuration 1 to 3 Seals (below): stiffness increases with thicker splined collar (4 & 5) C xx element of damping matrix increases with: smooth seals and adequate rotational speed

33 HPFTP Rotordynamic Development Effects of H/W modifications on shaft stability and frequency: attempts to add damping and stiffness at supports insufficient attempts to reduce cross-coupling of seals insufficient

34 HPFTP Rotordynamic Development Difficulties in predicting behavior

35 HPFTP Rotordynamic Development Conclusions: computer models strong indicators of trends: showed inadequacy of solving problem solely at bearing support showed desirability of using smooth seals showed sensitivity to nonlinearities could be fitted to engine runs, but never reliable in prediction of tests major destabilizers: seals and turbine cross-coupling other effects (deadband, impellers, internal hysteresis, etc.) secondary or unknown major stabilizers: increase of bearing stiffness increase of viscous damping of bearings (considerable increase necessary) bearing support stiffness asymmetry a stabilizer, high ratios needed ( 2.5:1) increase of stiffness and damping of smooth seals (effective only above rpm) friction damping at supports either ineffectual or even destabilizing stiffness data reliable, seal dynamics more uncertain More work needed

36 HPFTP Turbine Gas Dynamics and Cooling Turbine cooling required for high speeds: strength of material dependent on temperature turbine gas path temperatures: ~ 815 C (1500 F) high pitchline velocity: ~ 480 m/s 1600 fps) turbine disc, bearings & housing require cooling Liquid hydrogen selected for coolant: excellent cooling capability available in pump pressures sufficient to provide positive flow: turbine inlet pressure: ~ 391 atm (5750 psi) 1st nozzle discharge pressure: ~ 344 atm 5050 psi) Flow circuit designed to achieve cooling: operational at ~ 1.2 seconds after start: at ~ 7000 rpm lift-off seal opens ignition in preburner ~ 1,0 second all components cooled before high speed achieved

37 HPFTP Turbine Cooling Problems Inadequate LH2 coolant flow to bearings: results in hot gas backflow turbine bearings failed verified pressure in bearing cavity low

38 HPFTP Turbine Housing Problems Housing cracks and bulging experienced: manifested in numerous areas of turbine Differential rubbing noted on dynamic seals: alignment similar to that of housing failures Evidence of anomalously high temperatures: indication of unaccounted for hot gas flows Problems related to transverse Δp: generated by hot gas manifold design Turbine discharge-hot gas manifold (HGM) problems: flow exits turbine at high velocity exit guide vane flow separation tight turnaround duct leads to inner wall separation transverse Δp and flow velocity due to HGM design flow separation on some axial turbine struts nonuniform flow in HGM