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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
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
TS diagram of ideal vapor compression cycle
Satu
rate
d liq
uid
Saturate
d vapor
The basic vapor compression cycle
The ideal vapor compression cycle has two inherent cycle losses (thermodynamic irreversibilities): Throttling and desuperheating
The basic vapor compression cycle
Throttling loss
Tem
pera
ture
, TTe
mpe
ratu
re, T
Entropy, S
Vapor
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
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
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
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
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
Superheat loss
TS diagram of ideal vapor compression cycle
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
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
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
∆∆∆∆hevap,simple
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
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
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
∆∆∆∆hevap,simple
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
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
Superheat loss
TS diagram of the two-stage economizedideal vapor compression cycle
Satu
rate
d liq
uid
Saturated vapor
isenthalpic expansion
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
isentropic compression
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
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
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
isentropic compression
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
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
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
Heat In from air conditioned space
Bi-phaseturbine
Motor
Power In
GeneratorPower Out
∆∆∆∆hevap,simple
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
isentropic compression
Isobaric heat extraction
Isentropicexpansion
Liquid
Two-phase
Isobaric heat addition
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
Heat In from air conditioned space
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
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
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
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
Throttling loss
Tem
pera
ture
, TTe
mpe
ratu
re, T
Entropy, S
Vapor
isentropic compression
Isobaric heat extraction
isenthalpic expansion
Liquid
Two-phase
Isobaric heat addition
Superheat loss
TS diagram of ideal vapor compression cycle
Satu
rate
d liq
uid
Saturated vapor
Tem
pera
ture
, TTe
mpe
ratu
re, T
Entropy, S
Vapor
isentropic compression
Isobaric heat extraction
isentropic expansion
Liquid
Two-phase
Isobaric heat addition
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
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
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
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
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)
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)
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
Subcooled liquidregion
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
Subcooled liquidregion
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, ( )= −
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
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
Flow Fraction.25 .50 .75 1.0
.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,
ηη ηη/ ηη ηη
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,
ηη ηη/ ηη ηη
max
η ηturbine turbinex x x( ) ( ) ..
= = −0 01750175
TURBINE EFFICIENCY AS A FUNCTION OF ENTERING REFRIGERANT QUALITY
Comparison between test data and proposed correlation:
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
Comparison between test data and proposed correlation:
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
Flow Fraction.25 .50 .75 1.0
.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
Flow Fraction.25 .50 .75 1.0
.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
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
Flow Fraction.25 .50 .75 1.0
.8
.6
.4
Operation limit caused by compressor surgeDesign point
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%
•Introduction•Description of cycle modification•Description of turbine hardware•Control for off-design conditions•Performance at off-design conditions•Conclusions
Two-Phase-Flow Turbo-Expandersin Water-Cooled Chillers
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
Isobaric heat extraction
isenthalpicexpansion
Liquid
Two-phase
Isobaric heat addition
Superheat loss
TS diagram of ideal vapor compression cycle
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
Technology OverviewWhy are we pursuing this technology?
Tem
pera
ture
, TTe
mpe
ratu
re, T
Entropy, S
Vapor
isentropiccompression
Isobaric heat extraction
isentropicexpansion
Liquid
Two-phase
Isobaric heat addition
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
Technology OverviewWhy are we pursuing this technology?
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
∆∆∆∆hexpansion
Enthalpy, h
Pres
sure
, PPH diagram of the expressor test facility
LiquidTwo-phase
Vapor
Isoba
ric h
eat a
dditio
n
condensation
Adiabaticpressure rise
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
∆∆∆∆hexpansion
Enthalpy, h
Pres
sure
, PPH diagram of the expressor as intended
LiquidTwo-phase
Vapor
Adiabaticpressure rise
Com
pres
sion
∆∆∆∆hcompression
compcompandand hmhm ∆=∆ expexp
Expa
nsio
n
∆∆∆∆hexpansion
Enthalpy, h
Pres
sure
, PPH diagram of the expressor with
major liquid carry over
LiquidTwo-phase
Vapor
Adiabaticpressure rise
Com
pres
sion
∆∆∆∆hcompression
compcompandand hmhm ∆=∆ expexp
Expa
nsio
n
∆∆∆∆hexpansion
Enthalpy, h
Pres
sure
, P
LiquidTwo-phase
Vapor
Adiabaticpressure rise
Com
pres
sion
∆∆∆∆hcompression
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