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

3

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

4

« 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

5

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

⇒

6

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?

7

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

8

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

9

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

11

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

12

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 »

14

Three technological bricks