single vs two-stage (jekel and reindl 2008)

8/13/2019 Single vs Two-stage (Jekel and Reindl 2008)

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Todd B. Jekel, Ph.D., P.E., is assistant director and

Douglas T. Reindl, Ph.D., P.E., is professor and

director at the University of Wisconsin-Madison’s In-

dustrial Refrigeration Consortium in Madison, Wis.

About the Authors

4 6 A S H R A E J o u r n a l a s h r a e . o r g A u gu s t 2 0 0 8

Single- or

Two-Stage

CompressionBy Todd B. Jekel, Ph.D, P.E., Member ASHRAE; and Douglas T. Reindl, Ph.D., P.E., Member ASHRAE

Historically, the move toward multistage compression systems in industrial

refrigeration applications was rooted in demand for lower operating tem-

peratures. First generation compression technologies consisted of reciprocating

compressors, but later rotary vane machines also were used. These two machines

faced some key constraints in their operation at low temperatures. Reciprocating

and rotary vane compressors have physical compression ratio (ratio of absolute

discharge to absolute suction pressure) limits that are quickly approached as

operating temperatures decrease. Additionally, compressors and the oil used in

the compressors have limits on discharge gas temperature.

Most reciprocating compressors arelimited to compression ratios on the orderof 8:1. A common design saturated con-densing temperature for an ammonia sys-

tem is 95°F (35°C), which correspondsto a saturation pressure of 196 psia (1351kPa). This fixes the high-pressure sideof the compressor. The design low-sidepressure of the compressor would be es-tablished to avoid exceeding the compres-sion ratio limit of 8:1. With a discharge

The following article was published in ASHRAE Journal, August 2008. ©Copyright 2008 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in

paper form without permission of ASHRAE.




sion ratio increases (suction pressure decreasing), the advantagesof two-stages of compression and liquid expansion increases. Two key points are identified inFigure 2 . The point designated

byj indicates the point (– 18.5°F [– 25°C]) where the efficiencyof the two-stage system (Figure 1c ) is 10% better (lower hp/ton)

than a single-stage system (Figure 1a ) whilek

indicates the point(– 32°F [– 35.6°C]) where the two-stage system is 10% betterthan the single-stage system with two-stages of liquid expansion(Figure 1b ). It is also noteworthy that the efficiency advantage ofthe two-stage compression increases as the suction temperaturedecreases. At a saturated suction temperature of – 12°F (– 24°C)(8 psig [0.6 bar]) the two-stage and single-stage system with two-stages of liquid expansion have equal operating efficiency. The advantage of a two-stage compression system is improved

system efficiency, especially for lower temperature process re-quirements. The efficiency advantage does come with a price:more compressors. Figure 3 shows the total compressor size(i.e., volume flow rate, cfm) per ton of low temperature refrig-eration load. Across the range of saturated suction temperatures(– 45°F to – 5°F [– 43°C to – 21°C]), the two-stage system has anincremental compressor size of 1.5 cfm/ton to 2 cfm/ton (0.20L/[s · kW] to (0.27 L/[s · kW]) compared to a single-stage systemand 2.5 cfm/ton (0.34 L/[s · kW]) compared to a single-stagesystem with two-stages of liquid expansion. More compressorsmean more maintenance, more machinery room floor space,etc. The two-stage system will result in lower compression ratiooperation on each individual compressor which reduces wearand tear compared to a single stage system. These trade-offsare relatively complex and plant-dependent; however, they willaffect the total life-cycle cost of the system.

As a Function of Intermediate Pressure

In the context of two-stage compression systems, the con-cept of an “optimum” interstage pressure arises. An optimuminterstage pressure is one that will minimize the total powerconsumption for the system. A frequently applied approxima-tion for this optimum intermediate pressure is given by:1

,

·opt,int discharge,sat suction sat

P PP= (1)

where P opt,int represents an estimate of the optimum interme-diate pressure, P suction,sat is the absolute saturation pressure

corresponding to the suction conditions, andP discharge,sat is theabsolute saturation pressure corresponding to the discharge orcondensing conditions. This relationship equally divides thecompression ratio between the low-stage and high-stage of atwo-stage compression system. The actual optimum will dependon the operating efficiencies of individual compressors attachedto each suction pressure level. In addition, there are other factorsthat may restrict the ability to vary the intermediate pressure suchas temperature requirements for loads served by the intermedi-ate pressure level of the system. Free from any constraints onintermediate pressure,Figures 4 and5 show the effect of inter-mediate pressure on the efficiencies of the three systems for a

fixed saturated suction temperature of – 40°F (– 40°C) and – 25°F

Fi gure 1: Schematics of r efr igeration systems evaluated. 1a (top):

Single-stage; 1b (center) : Single-stage with two-stage liqui d expan-

sion; 1c (bottom): Two-stage.

Evaporative

Condenser(s)

High PressureReceiver Low Temperature

Recirculator

Low Temperature

Evaporator(s)

Compressor(s)

Evaporative

Condenser(s)

HighPressureReceiver

LowTemperature

Evaporator(s)

LowTemperatureRecirculator

MediumTemperatureRecirculator/

Intercooler

MediumTemperatureEvaporator(s)

BoosterCompressor(s)

High StageCompressor(s)

A

C

Evaporative

Condenser(s)

HighPressureReceiver Low

TemperatureEvaporator(s)

LowTemperatureRecirculator

MediumTemperatureRecirculator

MediumTemperatureEvaporator(s)

Low TemperatureCompressor(s)

Medium TemperatureCompressor(s)

B



Adverti sement formerl y in this space.




(– 32°C), respectively. The condensing condition is fixed at 85°F(29°C) saturated which corresponds to 150 psig (10.3 bar). The symbol indicates the theoretical optimum intermediate pres-sure for a two-stage compression system. Notice that in this case,the actual optimum (i.e., lowest hp/ton) occurs at intermediatepressures slightly higher than the theoretical optimum. Also notethat the optimum is fairly broad, that is, the intermediate pres-sure does not have a large effect on overall system efficiency. Toillustrate that point, the results shown inFigure 4 reveal that any

intermediate pressure between 27 psig to 50 psig (1.9 bar to 3.4bar) yields system efficiencies within 2% of the optimum.

In reality, most refrigeration systems do not have only lowtemperature loads. Nearly all plants will have refrigerationloads that demand higher temperature refrigerant: cooler spaces,bulk product storage tanks, production air-conditioning, post-pasteurization cooling, etc. The presence of these higher tem-perature loads will not significantly degrade the efficiency ofmeeting the low-stage loads due to the relatively broad optimumof the intermediate pressure. This is excellent news because thehigh-stage load temperature requirements can now dictate theintermediate pressure without compromising the efficiency of

meeting the low-stage loads.

As a Function of Condensing Pressure

The last factor considered is the condensing pressure. Fig- ure 6 shows the effect of saturated condensing temperature(pressure) on the efficiencies of the three systems for a fixedintermediate pressure of 30 psig (2 bar) (16.5°F saturated[– 8.6°C]) and – 25°F (– 32°C) saturated suction (1.2 psig [0.08bar]). As the condensing pressure is reduced, the compressionratio decreases and at approximately 64°F (18°C) saturated(100 psig [6.9 bar]) the efficiency of the two-stage system isthe same as for a single-stage compression system with two

stages of liquid expansion.

Conclusions The use of two stages of compression is common in low-tem-

perature industrial refrigeration systems. The efficiency benefit ofa two-stage system increases as the temperature requirements arelowered and is weakly dependent on the intermediate pressure. Thedecision on whether to configure the system for multiple-stagesof compression is one that should weigh both the advantages anddisadvantages at the suction pressures required by the loads. Asa general rule, two-stage compression systems should always beconsidered when loads demand low suction temperatures, particu-larly those lower than – 25°F (– 32°C) (1.2 psig [0.08 bar]).

Regardless of the number of stages of compression, configuring

the liquid side for two-stages of expansion will nearly always in-

Fi gure 2: Refr igeration system efficiency as a function of l ow-

temperature suction requirements.

Fi gure 3: Compressor di splacement per ton as a function of low-

temperature suction requirements.

F u

l l - L o a

d C o m p r e s s o r

E f fi c i e n c y

( h p

/ t o n

)

3.50

3.25

3.00

2.75

2.50

2.25

2.00

1.75

1.50

1.25– 55 – 45 – 35 – 25 – 15 – 5

Anhydrous Ammonia (R-717)

Twin Screw Compressor

External Oil Cooling

Intermediate Pressure = 30 psig (16.5°F sat)

Condensing Pressure = 150 psig (85°F sat)

Single-Stage Compression andTwo-Stage Liquid Expansion

Two-Stage Compression

and Liquid Expansion

Single-Stage

Compression

and Liquid

Expansion

Saturated Suction Temperature (°F)

j

k

T o

t a l C o m p r e s s o r

V o

l u m e

F l o w

R a

t e ( c f m / t o n

)

18

16

14

12

10

8

6

4– 55 – 45 – 35 – 25 – 15 – 5






Single-Stage

Compression and

Two-Stage Liquid

Expansion



Saturated Suction Temperature (°F)

Single-Stage Compression


Figure 4: Refr igerati on system efficiency (– 40°F [– 40°C] sat) as a

function of intermediate pressure.

F u

l l - L o a

d C o m p

r e s s o r

E f fi c

i e n c y

( h p

/ t o n

)

3.50

3.25

3.00

2.75

2.50

2.25

2.00

1.75

1.50

1.25




Suction Pressure = 9 in. Hg (– 40°F sat)


Single-Stage Compression and Liquid Expansion

Single-Stage Compression and Two-Stage Liquid Expansion

Two-Stage Compression and Liquid Expansion

Theoretical Optimum

Intermediate Pressure

10 15 20 25 30 35 40 45 50Intermediate Pressure (psig)



A u gus t 2 0 0 8 A SHR A E J our n a l 5 1

crease the efficiency of the refrigeration system. The only possibleexception would be a single-stage system with two temperature

levels during low condensing pressure operation with very smallcoincident intermediate pressure loads (e.g., winter operation incold climates). The condition results in a small amount of flashgas from liquid makeup to the low-temperature loads and oftenresults in low part-load efficiency of the intermediate pressurecompressor(s). In this case, it would be an advantage to be able toshut off the intermediate pressure compressors and provide liquidmakeup to the low-temperature recirculator package directly fromthe high pressure receiver during winter.

ReferencesStoecker, W.F. 1998.1. Industri al Refri gerati on Handbook. NY:

McGraw-Hill Publishers.



Figure 6: Refrigerati on system efficiency (– 25°F [– 32°C] sat) as a

function of condensing pressure.

F u

l l - L o a


E f fi c

i e n c y

( h p

/ t o n

)

3.50

3.25

3.00

2.75

2.50

2.25

2.00

1.75

1.50

1.2560 65 70 75 80 85 90 95 100




Suction Pressure = 1.2 psig (– 25°F sat)



Two-Stage Compression and Liquid Expansion


and Two-Stage LiquidExpansion

Saturated Condensing Temperature (°F)

Figure 5: Refr igerati on system efficiency (– 25°F [– 32°C] sat) as a

functi on of intermediate pressure.

F u

l l - L o a


E f fi c

i e n

c y

( h p

/ t o n

)

3.50

3.25

3.00

2.75

2.50

2.25

2.00

1.75

1.50

1.2515 20 25 30 35 40 45 50




Suction Pressure = 1.2 psig (– 25°F sat)





Theoretical Optimum

Intermediate Pressure


and Two-Stage Liquid

Expansion

Immediate Pressure (psig)

single vs two-stage (jekel and reindl 2008)

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