overall thermodynamic model of an ultra microturbine · overall thermodynamic model of an ultra...
TRANSCRIPT
Y. Ribaud,SFT, Mars06 1
OVERALL THERMODYNAMIC MODEL OF AN ULTRA
MICROTURBINE
• µturbine description
• separation between aero and heat losses
• « hot button » energetic model
• main results
•future research directions
2Stanford UniversityIHI, Sendaï, University
MIT
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burner compressor rotor vaned diffusor
Flame holders
NGV
Turbine rotor
MIT first model
1cm ×××× 3mm, 3 gr
Power : 10 - 50 W, 7 gr/h fuel
M air = 0.1g/s
N = 2.4 106 rpm
Tit = 1600 K
ΠΠΠΠ ∼∼∼∼ 3, pressure ratio
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« Hot button » thermodynamic model
• In tiny engines ⇑⇑⇑⇑ heat transfer by internal forced convection,
conduction in the material, radiation, external free convection.
• The compressor becomes a heat exchanger/compressor
• The turbine becomes a heat exchanger /turbine
• convection internal heat transfer : αααα(Re)*(Tw - Tg)* Sheat
• to compare with mechanical power ρρρρ.SinletV.∆∆∆∆Hi
• α ⇑⇑⇑⇑ for low Re, (Tw - Tg) important, S too important , Ns non optimal.
• With SiC, the Biot number , of the order of 0.1 : Bi = αααα L/λλλλw
Hot
Bu
tton
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Aerodynamic polytropic efficiency
Compressor case
aerodynamic losses : dPcomp.(1-ηp) = δδδδf+
ηp by definition is the aerodynamic polytropic efficiency
heat transfer calibration : λ = / Pcomp
0≠Q
Q
0=Q
( )1
2
1
2 .
1
Pi
Pi
Ti
Tipol
−
= ηγ
γ
( )1
2 .
1
2 )1(
1
Pi
Pi
Ti
Ti pol
+−
= ληγγ
⇒
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Heat transfer scheme and main calculation sections
stator parts
rotor parts
1
2
3
4
7
8
6
9
10
5
11
Heat exchange
12
Combustion
5’
13
a
g f
e
d c b
h i
e’
Balance in the stator
Balance in the rotor(s), option : conduction in the shaft
Conservation equations in the fluid volumes :
in the compressor rotor
in the vaneless diffusor
in the vaned diffusor
in the premixing region
in the burner
upstram the NGV
in the NGV
in the vaneless space between the NGV and the turbine rotor
in the turbine rotor
in the rotor/stator cavities
T stator ?
T rotor comp. ?
T rotor turbine?
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AERO and THERMODYNAMIC CALCULATIONS
Input :
• air mass flow rate mair,
• angular speed ω,
• aerodynamic polytropic efficiencies:
ηpc *ηpt=0.42
• compressor rotor work factor µ’∼ 0.8 = vtb/ωrb• T max. in the combustion chamber
• Tcombustor= Te’ ∼1600K for SiC.
• Microturbine geometry.
• Type of fuel : H2, kerosene, methane, propane
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Data input Exchange surface calculations ρρρρ, T, p, V calculations, adiabatic case
Thermodynamic balance, calculation of Ti,
Tstator, Troror, m(fuel)
Heat exchange coefficient calculations (h)
Heat transfer rate calculations Q
Compressor
pressure ratio
Turbine
expansion ratio
checkUpdate of Vt at the rotor
turbine inlet
Update aerothermal conditions ρρρρ, p, T, V Check the heat transfer coefficients Power calculations
no
no
Thermodynamic model
and organigram
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NET POWER (Watts) VERSUS AXIAL GAP
0
10
20
30
40
0 10 20 30 40 50 60 70
Axial gap of the electric generator (µm)
witho ut ins ula tio n
IGV ins ula tio n
Useful Power (W) Microturbine1cm
diameter
Microturbine2 cm diameter
Adiabatic 14 54
Non-adiabatic 4 24
Non-adiabatic+ disc friction
1.7 17
0
0.5
1
0.2 0.3 0.4 0.5
Useful power as function of the product of
the aero. efficiencies of the turbomachines
Useful power (2cm)
as function of T in combustor
0
10
20
12001300140015001600
Temperature (K)
Watts
Calculation gains
10
Net power for partial thermal insulation
05
1015202530354045
no in
sula
tion
burn
erdi
ffuso
rm
ixin
g ch
anne
ltu
rbin
e ro
tor
com
pres
sor r
otor
com
pres
sor
IGV
turb
ine
Watts
2 Temperatures 3 Temperatures
net power ( Watts), double µ turbine
18.3
19.7
17.7
17
17.5
18
18.5
19
19.5
20
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
stator emissivity
900
920
940
960
980
1000
T (K)
T_Stator (K) P Net (W)
Results (following)
28.4521.633.87
4.66E-06
Net Power (W)
Thermal Efficiency (%)Net Power with considering disk friction (W)
Fuel Mass Flow Rate (kg/s)
Hydrogen
25.2718.463.22
1.24E-05
Net Power (W )
Therm al Efficiency (%)Net Power with considering disk friction (W )
Fuel M ass Flow Rate (kg/s)
Propane
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57 W
convection 8 W
radiation 48 W
Thermal losses
in exhaust gases
( T = 1103 K )
454 W
Net Power
17 W
Work efficiency
3.2%
Air flow
0.4 g/sTair = 288 K
Tcompressor = 670 K
Tturbine= 840 K
Tchamber = 1600 KTstator = 930 K
Fuel : 46.6 g/h
7 W
19 W
34 W
16 W
12 W
23 W
5 W
34 W 81 W
28 W
9 W
51 W
4 W
PComb = 528 W
57 W
convection 8 W
radiation 48 W
Thermal losses
in exhaust gases
( T = 1103 K )
454 W
Net Power
17 W
Work efficiency
3.2%
Net Power
17 W
Work efficiency
3.2%
Air flow
0.4 g/s
Air flow
0.4 g/sTair = 288 K
Tcompressor = 670 K
Tturbine= 840 K
Tchamber = 1600 KTstator = 930 K
Fuel : 46.6 g/hFuel : 46.6 g/h
7 W
19 W
34 W
16 W
12 W
23 W
5 W
34 W 81 W
28 W
9 W
51 W
4 W
PComb = 528 W
Microturbine MEMS, 2cm
Static Pressure [Pa] 331431
Static Temperature [K] 1544
Velocity [m/s] 40
Static Pressure [Pa] 240794
Static Temperature [K] 1283
Velocity [m/s] 512
Condition Input
Inlet
Outlet
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Heat flux in the NGV for isothermal wall temperature T = 921K
Tatsuo ONISHI ASME RENO 05
COUSTEIX model 79 W
Conventional model 88W
average
Conventional model 72W
local
K. OKAMOTO model 112W
CFD by T. ONISHI 75 W
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CONCLUSION
Is the MEMS concept valid from an energetic point of view?
• Uncertainty about the aerodynamic compressor efficiency
• etching depth ? Depth variation with radius?
• Possibility of the NGV thermal insulation?
• Electric generator efficiency and thermal integration?
• Efficient minimum volume for the combustion chamber
• Gas bearings, : stability, lost power and stiffness
• new direction :
bulk machining µturbine, more than one material
« the die is better than the chip »
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Three technological bricks