Short Course on Throttle Loss Power Recovery in ... · Throttle Loss Power Recovery in...

Short Course on

Throttle Loss Power Recoveryin Refrigeration and Cryogenics

Joost J. BraszCarrier Corporation

Syracuse, New York 13221

21th IIR International Conference of RefrigerationWashington DC, August 17-22, 2003

Course Overview

• Section 1. Thermodynamics of Refrigeration and Cryogenic Cycles with Emphasis on Throttle Loss Power Recovery

• Section 2. Example of a Commercialized Throttle Loss Power Recovery System: Two-Phase-Flow Turbo-Expanders in Water-Cooled Chillers

• Section 3. Example of Throttle Loss Power Recovery R&D: The Expressor as Throttle Valve Replacement

SECTION 1Thermodynamics of Refrigeration and Cryogenic Cycles

with Emphasis on Throttle Loss Power RecoveryLiterature:1. Hays, L.G., History and Overview of Two-Phase Flow Turbines, Proceedings of the IMechE International Conference on

Compressors and their Systems, London, England, pp. 159-170. Sept 13-15, 1999.

2. Smith, I.K., Review of the Development of Two_phase Screw Expanders, Proceedings of the IMechE International Conference on Compressors and their Systems, London, England, pp. 95-104. Sept 13-15, 1999.

3. Granryd, E.G.U, Method of Improving Refrigeration Capacitry and Coefficient of Performance in a Refrigerating System, and a Refrigerating System for Carrying Out Said Method, US Patents 4,014,182, March 29, 1977

4. Zoughaib, A. and Clodic, D., A Turbo Expander Development for Domestic Refrigeration Appliances, 21th International Congress of Refrigeraytion 2003, Washington DC.

5. Heyl P. and Quack, H. Free Piston Expander-Compressor for CO2-Design, Applications and Results, Proceedings of the20th International Conference of Refrigeration, Sydney, Australia, 1999.

6. Bond, T., Replacement of Joule Thompson Valves by Two-Phase Flow Turbines in Industrial Refrigeration Applications, Proceedings of the IMechE International Conference on Compressors and their Systems, London, England, pp. 361-374. Sept 13-15, 1999.

The basic vapor compression cycle

Compressor

Evaporator/Cooler

Condenser

Vapor Compression CycleHeat Out to Ambient

Heat In from air conditioned space

Throttle Valve Motor

Power In

The basic vapor compression cycle used in the refrigeration industry consists of the following four processes:

1. Compression2. Condensation3. Throttling4. Evaporation

Thermodynamically, the vapor compression cycle is represented by a pressure enthalpy (PH) and/or a temperature entropy (TS) diagram.

∆∆∆∆hevap ∆∆∆∆hcomp

Enthalpy, h

Pres

sure

, P

PH diagram of the ideal vapor compression cycle

Liquid

Two-phase

Vapor

Isobaric heat rejection

Isobaric heat addition

isenthalpic expansion

isen

tropi

c

com

pres

sion

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor

isentropic compression

Isobaric heat rejection


Liquid

Two-phase


TS diagram of ideal vapor compression cycle

Satu

rate

d liq

uid

Saturate

d vapor


The ideal vapor compression cycle has two inherent cycle losses (thermodynamic irreversibilities): Throttling and desuperheating


Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor


Isobaric heat extraction


Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cycleindicating its two inherent irreversibilities

Satu

rate

d liq

uid

Saturated vapor

Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss

Satu

rate

d liq

uid

Saturated vapor

Throttle Loss (area B) is equal to area A on TS diagramArea A depends on slope of saturated liquid line and

increases when approaching the top of the dome

A

B

Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss


Satu

rate

d liq

uid

Saturated vapor

1

2

3

∫∫ +=−2

1

2

112 VdpTdshh ∫∫ +=−

3

1

3

113 VdpTdshh

Along isobar: ∫ = 0Vdp Also h2=h3 => Area12561=Area13461

456

7

=> Area127=Area3457

Throttle Loss Reduction Methods

1. Subcooling2. Multistaging with economizers3. Two-Phase Turbines4. Granryd Cycle

∆∆∆∆hevap,simple

Enthalpy

Pres

sure

D

A B

C

Refri

g. S

at. L

iqui

d

Refri

g. S

at. V

apor

∆∆∆∆hevap,sens. subc.

E

A

CYCLE IMPROVEMENT FROM SENSIBLE SUBCOOLING

Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cycle with sensible subcooling

Satu

rate

d liq

uid

Saturated vapor

subcooling

Conclusion: less throttling loss with subcooling


Enthalpy

Pres

sure

B

Refri

g. S

at. L

iqui

d

Refri

g. S

at. V

apor

∆∆∆∆hevap,flash subc.

A

EF

CYCLE IMPROVEMENT FROM “FLASH”SUBCOOLING

CD

A

Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cycle with flash subcooling

Satu

rate

d liq

uid

Saturated vapor

flashsubcooling

Conclusion: less throttling loss with sensible subcooling


Enthalpy

Pres

sure

D

A B

C

Refri

g. S

at. L

iqui

d

Refri

g. S

at. V

apor

∆∆∆∆hevap,economized

A

EF

CYCLE IMPROVEMENT FROM ECONOMIZING

Throttling losses

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss

TS diagram of the two-stage economizedideal vapor compression cycle

Satu

rate

d liq

uid

Saturated vapor


Conclusion: Throttling loss is half of that of simple cycle

Improvement of COP with increased number of stages and economizers

7.4

7.6

7.8

8

8.2

8.4

8.6

8.8

9

9.2

0 1 2 3 4 5Number of stages

Coe

ffici

ent o

f Per

form

ance

R134aR11R22Carnot

COP of ideal R134a system with and without subcooling and actual system with 80% efficient compressor

Tevap = 5 0C, Tcond=36 0C

5.5

6.0

6.5

7.0

7.5

8.0

8.5

9.0

0 1 2 3 4 5Number of stages

Coe

ffici

ent o

f Per

form

ance

no subcooling4 K subcooling4 K subcooling and Effcomp=.8

B Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor


Isobaric heat ex tracti on

isenthalpic expansi on

Liquid

Two-phase

Isobaric heat a ddi tion

Superheat loss

Throttle loss (area B on TS diagram) = area A

Satu

rate

d liq

uid

Saturate

d vapor

A



Conclusion: Cycle Efficiency less for higher operating T

Ideal vapor compression cycle COP of R11 and R134a for different evaporation temperatures at constant Carnot

efficiency (Tevap/(Tcond-Tevap)=8.973

6.0

6.5

7.0

7.5

8.0

8.5

9.0

9.5

200 250 300 350 400Evaporation temperature, K

Coe

ffic

ient

of P

erfo

rman

ce

R11R134aCarnot

R11 and R134a vapor compression cycle effectiveness:COPvapor compression cycle/COPCarnot cycle for different evaporation

temperatures at (Tevap/(Tcond-Tevap)=8.973

0.7

0.8

0.8

0.9

0.9

1.0

200 220 240 260 280 300 320 340 360 380 400

Evaporation temperature, K

Vapo

r co

mpr

essi

on c

ycle

effe

ctiv

enes

s

R11R134a

Conclusion: Cycle efficiency higher than component efficiencies

B Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor


Isobaric heat ex tracti on

isenthalpic expansi on

Liquid

Two-phase

Isobaric heat a ddi tion

Superheat loss

Throttle loss (area B on TS diagram) = area A

Satu

rate

d liq

uid

Saturate

d vapor

A



Conclusion: Cycle Efficiency less for refrigerants with lower Tcr

Tcr

Coefficient of Performance of Ideal Simple Cycle with 4K Subcooling of Different Refrigerants plotted against

Critical Temperature

7.7

7.8

7.9

8

8.1

8.2

8.3

8.4

8.5

350 375 400 425 450 475 500Critical Temperature, K

CO

P

R11 R141b

R134a

R245fa

R123

R12

R236fa

R114

R22

R113

Tevap = 5 0CTcond = 36 0CTsubc = 32 0CTsup = 5 0C

Compressor

Evaporator/Cooler

Condenser

Vapor Compression Cycle with Throttle Loss Power Recovery

Heat Out to Ambient


Bi-phaseturbine

Motor

Power In

GeneratorPower Out


Enthalpy

Pres

sure

D

A B

C

Refri

g. S

at. L

iqui

d

∆∆∆∆hevap,throttle loss power recovery

A

∆∆∆∆h expander

CYCLE IMPROVEMENT FROM THROTTLE LOSSPOWER RECOVERY (TLPR)

Throttling

Expansion

Refri

g. S

at. V

apor

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor



Isentropicexpansion

Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cyclewith throttle loss power recovery

Satu

rate

d liq

uid

Saturated vapor

Conclusion: All refrigerants have same efficiency with perfect throttle loss recovery

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor

LiquidTwo-phase

Superheat loss

TS diagram of the Granryd cycle

Satu

rate

d liq

uid

Saturated vapor

Compressor

Evaporator/Cooler

Condenser

Alternative method of throttle loss power recovery:Granryd Refrigeration CycleHeat Out to Ambient


Motor

Power In

Replacing the Joule-Thompson throttling valves in cryogenicequipment will lead to substantial energy savings and yield increases

Conclusions

Two-phase flow throttle loss recovery in air-conditioning and refrigeration equipment has the potential of improving systemoverall efficiency by 5-7 percent.

The potential of this technology for cryogenics seems to behigher in terms of yield increase and energy savings.

SECTION 2Example of commercialized throttle loss power recovery:

Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers

Literature:1. Brasz, J.J, , 1995, Improving the Refrigeration Cycle with Turbo Expanders, Proceedings of the 19th International

Conference of Refrigeration, Volume IIIa, pp 246-253, 1995.

2. Hays, L.G and Brasz, J.J., Two-Phase Turbines as Stand-Alone Throttle Replacement Units in Large 2000- 5000 Ton Centrifugal Chiller Installations, Proceedings of the 1998 International Compressor Engineering Conference at Purdue, Vol II, pp. 797-802, 1998.

3. Brasz, J.J. Performance Characteristics of Two-Phase Flow Turbo-Expanders used in Water-Cooled Chillers, Proceedings of the IMechE International Conference on Compressors and their Systems, London, England, pp. 171-180. Sept 13-15, 1999.

4. Brasz, J.J. Two-Phase Flow Turbine, US patent 5,467,613, November 21, 1995

•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions




CFC11, the most efficient refrigerant CFC11, was banned in 1995

The second best refrigerant, HCFC123, will be banned in the near future

HFC134a suffers 5.5% cycle efficiency penalty relative to CFC11 and a4.5% penalty relative to HCFC123

The bi-phase turbine was introduced to overcome the cycle disadvantagesof chlorine-free refrigerants

The bi-phase turbine is currently a standard option on HFC134a chillers

The focus of this paper is bi-phase turbine performance at off-designconditions

Introduction



Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor




Liquid

Two-phase


Superheat loss


Satu

rate

d liq

uid

Saturated vapor

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor



isentropic expansion

Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cyclewith throttle loss recovery

Expansion Throttling

∆∆∆∆hevap,throttle loss power recovery∆∆∆∆hturbine

∆∆∆∆hevap,simple ∆∆∆∆hcomp

−−−− ∆∆∆∆hturbine

Enthalpy, h

Pres

sure

, PPH diagram of vapor compression cyclewith and without throttle loss recovery

Liquid

Two-phase

Vapor



CONVENTIONAL COMPRESSOR

TURBO-ASSISTED COMPRESSOR

CONVENTIONAL COMPRESSORTURBINE

Nozzle with inlet flow divider

Nozzle Block

TURBINE WHEEL

Cut-away view of the integrated two-phase-flow turbineused on centrifugal chillers in the

1000 to 2000 kW cooling capacity range

Turbo-assisted Centrifugal Chiller



TURBO-ASSISTED COMPRESSOR

Rotating Inlet Guide Vanes for Capacity ControlNo Variable Geometry

TURBINE COMPRESSOR

Hea

d Fr

actio

n

Flow Fraction.25 .50 .75 1.0 1.25

.8

.6

.4

Design point

.2

1.

.0.0

Inlet guide vane capacity control as function of head and flow

IGV

=90 degrees

IGV

=70 degrees

IGV

=50 degrees

IGV

=30 degrees

IGV

=20 degrees

IGV

=10 degrees

Operation limit caused by compressor surge

1.2

CROSS-SECTION OF A TWO-PHASE-FLOW NOZZLE

Liquid saturation line

Throttling(H=constant)

Subcooled liquidregion

Enthalpy, h

Pres

sure

, PPr

essu

re, P

Two-phaseregion

PH diagram comparingthrottling valve (isenthalpic expansion) to

power recovery turbine (isentropic expansion)

Power recovery(S=constant)

Observations:

1. No difference between isentropicand isenthalpic expansion insubcooled liquid area

2. Not much difference betweenisentropic and isenthalpic expansionduring initial expansion in two-phaseregion

Liquid saturation line

Throttling(H=constant)


Enthalpy, h

Pres

sure

, PPr

essu

re, P

Two-phaseregion

PH diagram comparingthrottling valve (isenthalpic expansion) to

power recovery turbine (isentropic expansion)

Power recovery(S=constant)

V P Pliq

21 22= −( )

.ρ

V h h3 1 32= −( )

V1 0=1

2

Liqu

id s

atur

atio

n lin

e

Two-phaseregion


Trajectory of nozzle expansion process inpressure-enthalpy diagram

Pres

sure

, PPr

essu

re, P

Enthalpy, h

3

V P Psat liq

liq2

1 22, . .

.

( )= −ρ

V h hspouting3 1 32, ( )= −

Nozzle capacity controlled byliquid velocity at throat:

13

2

Nozzle kinetic energy controlledby exit velocity:

CROSS-SECTION OF A TWO-PHASE-FLOW NOZZLE

1

2

Liqu

id s

atur

atio

n lin

e

Two-phaseregion


Trajectory of nozzle expansion process inpressure-enthalpy diagram

Pres

sure

, PPr

essu

re, P

Enthalpy, h

3

Capacity is controlled byamount of subcooling

Spouting velocity is controlled by enthalpy drop

These two relationshipsare virtually uncoupledsince enthalpy does not change during throttlingof saturated liquid.

V P Psat liq

liq2

1 22, . .

.

( )= −ρ

V h hspouting3 1 32, ( )= −



Hea

d Fr

actio

n

Flow Fraction.25 .50 .75 1.0

.8

.6

.4

Turbine operation limit caused by compressor surgeDesign point

.2

1.

.0.0

Problem statement

Turb

ine

oper

atio

n lim

it ca

used

by

com

pres

sor c

hoke

How does turbine efficiency change with conditions ?

100%90%

80%70%

60%Hea

d Fr

actio

n


.8

.6

.4

Operation limit caused by compressor surgeDesign point

40%30%

20%10%

50%

.2

1.

.0.0

Relative efficiencies (ηηηη/ηηηηmax) of a centrifugal compressor withinlet guide vane capacity control as function of head and flow

Ope

ratio

n lim

it ca

used

by

com

pres

sor c

hoke

0.60

0.65

0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.0 1.5 2.0 2.5 3.0 3.5

Rel

ativ

e ef

ficie

ncy,

ηη ηη/ ηη ηη

max

Ratio of spouting velocity and rotor speed, Vspouting/Urotor

Relative turbine efficiency as a function of speed ratiofor different levels of inlet subcooling

5K subcooling

2K subcooling

1K subcooling

Rel

ativ

e ef

ficie

ncy,


max

Ratio of spouting velocity and rotor speed, Vspouting/Urotor

RELATIVE TURBINE EFFICIENCY AS A FUNCTION OFSPEED RATIO AND INLET SUBCOOLING

Comparison between test data and proposed correlation:

0.60

0.65

0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1 1.5 2 2.5 3 3.5

1K subcooling2K subcooling

5K subcooling

η ηturbineis

rotorsubc

VU

T= + − − −0 5 1 122 5 1296 8 33 1752. [ sin( . . *( . ) )]max ∆

Turbine inlet quality x=(hin-hliq)/(hvap-hliq)

Rel

ativ

e ef

ficie

ncy,


max

η ηturbine turbinex x x( ) ( ) ..

= = −0 01750175

TURBINE EFFICIENCY AS A FUNCTION OF ENTERING REFRIGERANT QUALITY


0

0.2

0.4

0.6

0.8

1

1.2

0 0.03 0.06 0.09 0.12 0.15

test datacorrelation

Mass flow rate correlation forsubcooled liquid entering the turbine

m A T P P Tturb throat liquid turb in turb in sat turb in= −2ρ ( ) *[ ( )], , ,

m A T P P Tturbine throat liquid turbine in turbine in sat turbine in= − −2 2 2ρ ( )*[ ( . )], , ,

Mass flow rate was found to be substantially larger than correspondsto maximum pure liquid velocity in nozzle throat:

An accurate mass flow rate prediction (within 5%) was possible byassuming a delay of vaporization of 2.2 K:

m x x x m xturbine turbine( ) ( . ) * ( )= − + =16 5 2 1 02

TURBINE FLOW RATE AS A FUNCTION OFENTERING REFRIGERANT QUALITY


0.5

0.6

0.7

0.8

0.9

1

0 0.03 0.06 0.09 0.12 0.15

test datacorrelation

Turbine inlet quality x=(hin-hliq)/(hvap-hliq)

Frac

tiona

l mas

s flo

w ra

te, m

(x)/m

(x=0

)

Using these four correlations

a turbine performance map could be developed

η ηturbineis

rotorsubc

VU

T= + − − −0 5 1 122 5 1296 8 33 1752. [ sin( . . *( . ) )]max ∆

η ηturbine turbinex x x( ) ( ) ..

= = −0 01750175

m A T P P Tturbine throat liquid turbine in turbine in sat turbine in= − −2 2 2ρ ( )*[ ( . )], , ,

m x x x m xturbine turbine( ) ( . ) * ( )= − + =16 5 2 1 02

100%

90%

80%70%60%

Hea

d Fr

actio

n


.8

.6

.4

Turbine operation limit caused by compressor surgeDesign point

40%30%20%10%

50%

.2

1.

.0.0

Relative efficiencies (ηηηη/ηηηηmax) of two-phase flow turbinewith inlet throttle capacity control as function of head and flow

Turb

ine

oper

atio

n lim

it ca

used

by

com

pres

sor c

hoke

0%

Spouting velocity < rotor speed

100%90%

80%70%

60%Hea

d Fr

actio

n


.8

.6

.4


40%30%

20%10%

50%

.2

1.

.0.0

Relative efficiencies (ηηηη/ηηηηmax) of a centrifugal compressor withinlet guide vane capacity control as function of head and flow

IGV

=90 degr ees

IGV

=70 degr ees

IGV

=50 degrees

IGV

=30 degrees

IGV

=20 degrees

IGV

=10 degrees

100%

80%

60%

Hea

d Fr

actio

n


.8

.6

.4


40%

.2

1.

.0.0

Comparison of turbine and compressor relative efficiencies(ηηηη/ηηηηmax) as function of head and flow

Ope

ratio

n lim

it ca

used

by

com

pres

sor c

hoke100%

80%

60%

40%



Conclusions

Two-phase-flow turbines applied to water-ccoled chillers are efficientlycontrolled by a float valve at the inlet of the turbine.

Turbine efficiency does not deteriorate with reduced mass flow rates as long as the input head stays constant and the fluid entering the turbine remains in the liquid phase.

A reduction in input head by 75% is required to reduce the spouting velocity to the rotor speed at which point the turbine becomes ineffective. Those conditions are unlikely to occur in practice.

The relative part-load efficiency of the bi-phase turbine is better than thatof a centrifugal compressor with variable inlet guide vanes.

Literature:1. Brasz, J.J, Smith, I.K. and Stosic, N., Development of a twin screw expressor as a throttle valve replacement

for water-cooled chillers, Proceedings of the 2000 International Compressor Engineering Conference at Purdue, Volume II, pp.979-986, Purdue University, West Lafayette, IN

2. Brasz, J.J., Single rotor pair expressor as two-phase flow throttle valve replacement, US patent 6,185,956 B1, February 13, 2001.

3. Brasz, J.J. Screw-expressor testing on an R-134a chiller: efficiency, liquid carry-over and chiller benefit, IMechE Conference Compressors and their Systems, London, England, September 2003.

SECTION 3Example of throttle loss power recovery R&D:The expressor as throttle valve replacement

Technology OverviewWhy are we pursuing this technology?

Throttling loss

Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor

isentropiccompression


isenthalpicexpansion

Liquid

Two-phase


Superheat loss


Satu

rate

d liq

uid

Saturated vapor

• Throttling loss is the only major fundamental loss of the vaporcompression cycle when using fluorocarbon refrigerants

• Difference in cycle efficiency between refrigerants is due to difference in throttling loss


Tem

pera

ture

, TTe

mpe

ratu

re, T

Entropy, S

Vapor

isentropiccompression


isentropicexpansion

Liquid

Two-phase


Superheat loss

TS diagram of ideal vapor compression cyclewith throttle loss recovery

The vapor compression cycle with throttle loss recovery approaches the maximum obtainable efficiency Carnot cycle


Throttle loss recovery is the only way to overcome the thermodynamic cycle disadvantage of R-134 versus R-123 and R-245fa

Cycle efficiency comparison different refrigerants (Assumptions: single stage compressor at following conditions:

42/96/0/7, ηηηηcomp=82%)

7.137.04

6.92

6.72

7.24

6.4

6.5

6.6

6.7

6.86.9

7

7.1

7.2

7.3

CFC11 HCFC123 HFC245fa HFC134a HFC134a +70% eff Exp

CO

P

Existing Throttle Loss Recovery: 19XRT

Turbo-assisted Centrifugal ChillerTURBO-ASSISTED COMPRESSOR

CONVENTIONAL COMPRESSORTURBINE

Two-phase flow turbine issues• Turbine application is expensive. It requires

– low-loss sensible subcooler– two throttle valves (upstream and bypass)– low-loss large piping from bottom condenser to compressor and back to

bottom evaporator– low-loss refrigerant distribution piping inside cooler

• Turbine-assisted concept with turbine attached to compressor drive limits its capacity range and prevents usage for other compressor concepts

• Turbine peak efficiency 52%

• Result: turbine only viable as a option on a limited number of chillers, not as a standard feature

Alternative two-phase flow expansion technology:

screw expanders

ADVANTAGES: • Expansion efficiencies up to 70% reported• Easier to apply to larger and smaller compressors

(not limited by the head/capacity/speed relationship of a turbo-machine)

Cross-section of two-phase flow screw expander

Expa

nsio

n

Thro

tt lin

g

∆∆∆∆hexpansion

Enthalpy, h

Pres

sure

, PPH diagram of the screw expressor test facility

LiquidTwo-phase

Isoba

ric h

eat a

dditio

n

condensation

Adiabaticpressure rise

R113 two-phase flow expander test facility

Photograph of two-phase-flow R113 screw expander at test facility

Online terminal display of test results

Screw expander efficiency as a function of speed

0102030405060708090

100

0 1000 2000 3000 4000 5000

Male rotor speed, rpm

Adi

abat

ic e

ffici

ency

COST PROBLEMS OF SCREW EXPANDERS

• If applied like the turbine (attached to the extension of the motor of the main compressor ) the applied screw expander cost looks even worse than applied turbine cost

• If recovered power is fed to a separate generator the efficiency advantage of the screw disappears due to electrical power conversion losses and the applied cost increases again

Maincomp-ressor

DriveMotor

Evaporator

Condenser

Expre ssor

Vapor compressed by expressor

Possible solution to the cost problem of throttle loss power recovery: the

EXPRESSOR

169

mm

280 mm465 mm

Dimensions of the expressor for a 500 ton chiller

18.3 inches

11.0 inches

6.7

inch

es

Single rotor pair expressor

Expa

nsio

n

Thr o

ttlin

g


Enthalpy, h

Pres

sure

, PPH diagram of the expressor test facility

LiquidTwo-phase

Vapor

Isoba

ric h

eat a

dditio

n

condensation


Com

pres

sion

∆∆∆∆hcompression

Thro

t t lin

g

compcompandand hmhm ∆=∆ expexp

Two-phase flow R113 expander/expressor test facility

How to determine expressor efficiency?

ηηη CompExpandExpressor . = Definition of expressor efficiency:

Where

h.mh.m =

CompComp

comp isentCompComp ∆

∆η

h.mh.m =

expand isentExpand

ExpandExpandExpand ∆

∆η

h.mh.m

h.mh.m =

expand isentExpand

ExpandExpand

CompComp

comp isentCompExpressor ∆

∆∆

∆*η

Substitution gives

hh.

mm =

expand isent

comp isent

Expand

CompExpress ∆

∆η

Therefore:

Two-phase flow expander/expressor test facility

mexpander

mcompressor

Measured expressor overall efficiencies using

Expressor overal efficiency from the flow measurement

0

10

20

30

40

50

60

70

80

90

100

0 2 4 6 8 10 12 14

Pressure p1 [bar]

eta

[%] First mach. Vc=1.308

First mach. Vc=1.081New mach., Vc=1.186

hh.

mm =

expand isent

comp isent

Expand

CompExpress ∆

∆η

Problem with experimental expressorefficiency determination

• Mass flow rate of vapor being compressed can not be measured when liquid carry-over takes place

Solution

• Assume a volumetric efficiency and determine mass flow rate from vapor density at compression inlet and voluymeflow rate as determined by the swept volume corrected with the volumetric efficiency:

ntdisplacemevolumetricvaporcomp Vm &⋅⋅= ηρ

Expressor efficiencies as measured at R113 test rig facility

Maincomp-ressor

DriveMotor

Evaporator

Condenser

Expre ssorVapor compressed

by expressor

high pressure liquid

low pressure liquid/vapor

Throttlevalve

high pressure vapor

low pressure vapor

P1=872 kPaT1=33.7 0C

P3=412 kPa

P5=826,862,886,937 kPa

P=887 kPa

P=362 kPa

Schematic of prototype expressor test set-up as throttle valve replacement at an R134a

chiller

R134a expressor testing at a chiller

Chiller with prototype expressor under construction

High pressure liquid

Expressor

High pressure vapor

Expander compressor

Low pressure two-phase flow

x2x4

x3

1 5

3

Expressor compressor inlet quality x4 versus expander exit quality x2

0%

5%

10%

15%

20%

25%

30%

1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000

expressor rpm

Qua

lity

at c

ompr

essi

on in

let

port

x4

x2

Liquid carry-over into vapor compressorsection of expressor

Liquid carry-over into vapor compressorsection of expressor

Expressor separation effectiveness of four test points close to design conditions

0%1%2%3%4%5%6%7%8%9%

10%

1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000

expressor rpm

Sepa

ratio

n ef

fect

iven

ess,

(x4-

x 2)/(

1-x 2

) 1

2

3 4

Expa

nsio

n


Enthalpy, h

Pres

sure

, PPH diagram of the expressor as intended

LiquidTwo-phase

Vapor


Com

pres

sion



Expa

nsio

n


Enthalpy, h

Pres

sure

, PPH diagram of the expressor with

major liquid carry over

LiquidTwo-phase

Vapor


Com

pres

sion



Expa

nsio

n


Enthalpy, h

Pres

sure

, P

LiquidTwo-phase

Vapor


Com

pres

sion


redu

ctio

n of

vap

or q

ualit

ydu

e to

exp

ress

or

throttle

3

xxBenefitSystem

−−=

11_

xthrottle

x3

x2

Prevention of liquid carry-over

Physical separation of expansionand compression section

Picture of prototype gate rotorswith slot to accommodatephysical separation of expansionand compression section

Compressor inlet quality x4 and expander exit quality x2 of original and baffled expressor

0%5%

10%15%20%25%30%35%40%45%50%

1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000

expressor rpm

Qua

lity

x4 original expressor

x4 baffled expressor

x2 baffled and original expressor

Expressor separation effectiveness of test points close to design conditions

0%5%

10%15%20%25%30%35%40%45%50%

1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000

expressor rpm

Sepa

ratio

n ef

fect

iven

ess,

(x4 -

x2)/(

1-x2

)

original expressor

baffled expressor

0%

10%

20%

30%

40%

50%

60%

70%

80%

90%

100%

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100%Separation effectiveness

Loss

in e

xpre

ssor

ben

efit

due

to

liqui

d ca

rry-

over

original expressor

baffled expressor

Expressor system benefit as a function of its separation effectiveness

Conclusions

The expressor concept has the potential of being a low cost throttle valve replacement

System benefit depends on the reduction in the two-phase flow quality entering the evaporator of the refrigeration system

Early prototypes suffered from liquid carry-over into the compressor section of the expressor

Liquid carry-over can be reduced by inserting a separation baffle plate inside the rotor

Date post:	22-Jul-2018
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