Todd Jekel, P.E., Ph.D. Assistant Director, IRC … flooded recirculation Evaporative Condenser(s)...

Energy efficient food processing:focus on refrigeration

Refrigeration Systems Review

Energy efficient food processing:focus on refrigeration

Refrigeration Systems Review

Todd Jekel, P.E., Ph.D.

Assistant Director, IRC

University of Wisconsin‐Madison

University of Wisconsin-Madison

Introduction

• Review of system types• Single stage

• direct‐expansion

• flooded

• overfeed

• Multi‐stage (compound)• direct

• indirect

• Cascade

• Energy efficiency• Pre‐requisites• Important considerations

• Keys to success

Evaporator configurations

• Gravity flooded

• Liquid overfeed• CPR‐fed• Recirculator‐fed

• Direct‐expansion (DX)

Direct‐eXpansion (DX)

Evaporator Configurations

• Air‐cooling

• Chiller

• Plate‐type

• Shell‐and‐tube

• Other

• Bulk silos

Liquid Feed Arrangements

• Thermostatic expansion valve

• Mechanical valve (Proportional control)

• Electronic

• Pulse‐width modulating (fast‐acting solenoid)

• Motorized modulating (continuous)

• Requires pressure & temperature transducer

• Requires controller

Single stage DX system ‐ traditional

EvaporativeCondenser(s)

Equalizeline

Un-protected suction

Compressor(s)

HighPressureReceiver

TT

High pressure liquid

High pressure gas

DX Evap 1

Protected suction

…

1

2

3

4

Solenoid valve Thermostaticexpansion valve

Equalizing line

T

DX Evap n

T

DX Evap n

King valve

RefrigerantTransfer System

To HPRSuction Trap

Single stage DX system ‐ emerging

Compressor(s)

Eq

ual

izer

line


High pressure gas

King valve (automatic)

CondenserEvaporativeCondenser


DX Evap 1

Protected suction

…

SuctionTrap

DX Evap n


To HPR

Dis

char

ge

line

CondenserEvaporativeCondenser

MotorizedValve

25 °F

TPController

Power

25 °F

TP

Direct‐eXpansion (DX) System

Evaporativecondensers

Dry suction

Compressor(s)


TT


DX Evap 1

Protected suction

…

To high pressure receiverSuction

trap

1

4


Equalizing line

T

DX Evap n

T

DX Evap n

King valve

High pressure gas

2

Typical conditions: 100 psi < P < 180 psi

130°F < T < 230°F

Typical conditions: 100 psi < P < 180 psi

56°F < T < 95°F

Low-side Condition Range: 24 psi < P < 75 psi

10°F < T < 50°F

Equalizerline

3


Compressor, rotary screwCompressor

Motor

Motor

Oil Separator

Oil Separator

9

Oil

Compressor discharge

MotorCompressor

1st Stage Oil Separation 2nd Stage Oil Separation

Dischargevapor

Oil Separator

Compressor, rotary screw

10

SuctionDischarge

Thermosiphon (refrigerant)oil cooling heat exchanger

Compressors, reciprocatingSuctionDischarge



Dry suction

Compressor(s)


TT


DX Evap 1

Protected suction

…


trap


Equalizing line

T

DX Evap n

T

DX Evap n

High pressure gasTo plant for evaporatordefrost


Condensers, evaporative

Evaporative condenser, HX

14



Dry suction

Compressor(s)


TT


DX Evap 1

“Wet” suction

…


trap


Equalizing line

T

DX Evap n

T

DX Evap n



Receivers, high pressure




Compressor(s)


TT


DX Evap 1

Protected suction

…


trap


Equalizing line

T

DX Evap n

T

DX Evap n


King valve


King valve




Compressor(s)


TT


DX Evap 1

Protected suction

…


trap


Equalizing line

T

DX Evap n

T

DX Evap n



Evaporator technologies

• Air‐cooling• Space conditioning higher temperature spaces(production or storage), coolers, holding freezers

• Ceiling‐hung or penthouse unit configurations

• Liquid‐cooling (secondary fluids and product)• Shell‐and‐tube

• Plate‐and‐frame

• Falling film

• Scraped surface

20

Evaporator, air‐cooling

Ceiling‐hung evaporator in a dock area

Penthouse evaporator in a freezer

temperature sensing bulb

external equalize line

TXVsolenoid

DX Fluid Product Silo




Compressor(s)


TT


DX Evap 1

Protected suction

…


trap


Equalizing line

T

DX Evap n

T

DX Evap n



Transfer System

Suction trap

Transfer drums

Oil pot

Ammonia DX evaporators

• Advantages

• Relatively low first cost

• Easy to build

• Minimal refrigerant inventory at the unit

• No “wet riser” issues to deal with

• With emerging technology, lower operating temperatures are achievable

• Allows use of EPRs

DX evaporators

• Disadvantages• Potential for liquid carryover to compressor

• Suction trap essential with transfer capability required

• Evaporator operating temperature limit (~10°F for ammonia)

• New technology is extending the operating temperature range

• Lower evaporator pressure required to achieve superheat• Other loads often dictate lower intermediate or high‐stage suction

• Electronic expansion valves mitigate but require more sensors and controls

• Refrigerant distribution problems at coil have to be managed• Flash gas & loss of liquid wetting due to stratification

• Head pressure requirements• Requires pressure differential for proper expansion valve function

Gravity flooded recirculation


Equalizerline



High pressure gas

Flooded evap 1

“Protected” suction

To HPR

King valve

Flooded evap, n

Dis

char

ge

line

Compressor(s)

SuctionTrap

Transfer Station

Solenoid valve

Flooded evaporator

Hand expansion valve

Float

1

2

3

4

surge drum

Evaporator, liquid supply

Evaporator, vapor return HPL

Strainer

Solenoid

HEV

Fill float

Gravity flooded air unit evaporator

Evaporators, liquid chillers

Plate‐and‐frame liquid chiller Shell‐and‐tube liquid chiller

Gravity flooded chiller evaporator

Plate heat exchan

ger

AB

CD

Flooded gravity recirculation system

• Advantages• good evaporator heat transfer characteristics

• simple evaporator operation (no pumps)

• easier to manage suction lines (no two phase flow)

• defrost condensate/return easier to manage

• can accommodate evaporator pressure regulators

• Disadvantages• oil management (each evaporator has to be drained)

• surge drum required for each evaporator• initial cost & on‐going mechanical integrity requirements

• high refrigerant inventory

• tends to have a large % of charge in production areas

Overfeed system layout

EvaporativeCondenserEvaporative

Condenser

Eq

ual

izer

line

Dry suction



High pressure gas

1

2

3

King valve

Pumpedrecirculator

4’Overfed evaporator(s)

4

T4”’

Wetreturn

3

EvaporativeCondenserEvaporative

Condenser

Dis

char

ge

line

Compressor(s)

Liquid overfeed system

Recirculator

Float column

Liquid refrigerant pumps

Pumped liquidline

Overfeed System

• Advantages• good evaporator heat transfer characteristics• ability to handle multiplicity of evaporators• allows for system expansion• eliminates need for multiple surge vessels• excellent part‐load or turn‐down capability• good compressor protection (from liquid)• ability to significantly float head pressure

• Disadvantages• system first cost• evaporator performance suffers when evaporator pressure regulators are used

• mechanical pump maintenance

Single stage compression, multiple temps


High pressure receiver

MediumTemperatureEvaporator(s)

Med

ium

Tem

per

atu

reR

ecir

cula

tor

High temperatureCompressor(s)

LowTemperatureEvaporator(s)

Lo

wT

emp

erat

ure

Rec

ircu

lato

r

Low TemperatureCompressor(s)

EqualizerEvaporative

Condenser(s)

LowtemperatureEvaporator(s)

IntercoolerLowtemperaturerecirculator

BoosterCompressor(s)

High-StageCompressor(s)


1

2

3

4

5

5’

6

4’5’’

King valve

Two‐stage compression(single temperature level direct intercooled with single stage liquid expansion)

High stage discharge gas line

Booster discharge gas line

Bo

ost

er s

uct

ion

Hig

h s

tag

e su

ctio

n

7

Two‐stage compression(single temperature level direct intercooled with two stages of liquid expansion)


High pressure receiver

MediumTemperatureEvaporator(s)

High-stageCompressor(s)

LowTemperatureEvaporator(s)

Lo

wT

emp

erat

ure

Rec

ircu

lato

r

BoosterCompressor(s)

Intercooler/MT Recirc

High stage discharge gas line

Recap

• Review of system types• single stage compression with evaporators configured as

• direct‐expansion

• flooded

• overfeed

• multi‐stage compression with liquid expansion configured as• direct

• indirect

38

Energy EfficiencyLow‐side Opportunities

Todd B. Jekel

Research Scientist

Industrial Refrigeration Consortium

Refrigeration efficiency

-30 -20 -10 0 10 20 3050

100

150

200

250

300

0.75

1

1.25

1.5

1.75

2

2.25

2.5

2.75

3

SST [F]

Capacity, tonsCapacity, tons

hphp

hp

/to

n

hp/tonhp/ton

Saturated Suction Temperature [°F] (i.e. pressure)

Compressor hp/ton

Compressor capacity or horsepower

Persistence of efficiency

• The efficiency gains of increasing the suction pressure are persistent

• Occur EVERY hour of operation• Not a function of ambient conditions (i.e. condensing pressure)

Low‐side Energy Conservation Measures (ECMs)

• Suction pressure set point changes• Raise suction pressure set point

• Add evaporator surface area• Go to first bullet

• Load reduction (infiltration, lighting, etc.)• Reduce suction piping pressure drop• Separate regulated loads ontonew suction level In

crea

sing level of difficulty

Suction Set Point Changes

• Operations are conservative by nature• Not uncommon to see suction pressure set points lower than required for loads

• Look first at all refrigeration load temperature requirements served by that suction pressure, then

• Look at evaporator liquid feed solenoid operation on the lowest temperature requirement to assess opportunity

Increasing Suction Pressure• Benefits

• reduced system energy use

• 1.5% reduction for each psi increase in suction pressure at Florida facility

• increased system capacity

• ~2% increase per psi increase in suction for typical high‐stage

• >7% increase per psi increase in suction for low‐stage (below 0 psig)

• prolonged compressor life & decreased oil cooling loads

• When

• all hours of the year

Evaporator ECMs

• Fan control• Duty cycling• VFD (more later)

• Defrost (more later)

• Liquid feed (overfeed systems)• Revisit, set, and log metering valve setting

• Maintenance• Maintain clean surfaces

Suction Set Pressure Change Constraints

• Compressor motor size

• Wait, you said the efficiency of the compressor goes up?

• Right, it does, but so does the capacity (i.e. mass flow rate). More capacity means more power.

• Compressor oil separator

• More mass flow rate means more velocity in the oil separator which means lower oil separation efficiency

• More on this later

Increase evaporator surface area• Effects

• Increases the required evaporating pressure to meet the same load (at the same temperature)

• Increases fan or pump power (parasitic)

• Considerations• Only add evaporator area on the refrigeration load with the lowest temperature requirement

• Let’s do an example…

Consider the opportunities

70°F space conditioning 65 psig

Water chiller 43 psig

40°F space conditioning 45 psig

35°F cooler 40 psig

Ice bank 25 psig

-25°F freezer

1.4 hp/ton @ 24 psig

{1.0 hp/ton @ 39 psig

ENERGY EFFICIENCYHIGH‐SIDE OPPORTUNITIES

Floating Condensing Pressure is refrigeration’s “Greatest Hit” of energy efficiency

• Everyone knows that reducing condensing pressure

• DECREASES refrigeration system operating cost

• DECREASING compressor operating cost, even though you are

• INCREASING evaporative condenser fan operating cost

December, 1910

Condensing Pressure ECMs

• Condensing pressure set point changes• Lower condensing pressure set point

• Add condenser surface area• Go to first bullet

Increa

sing level of difficulty

Condenser ECMs

• Fan/pump control• Fan VFD (more later)

• Maintenance• Maintain clean surfaces• Non‐condensables (purger)

Condensing pressure control

How do we control condensing pressure in industrial refrigeration systems?


• Our heat rejection system controls head pressure

• Increasing heat rejection rate causes head pressure to decrease

• Decreasing heat rejection rate causes head pressure to increase

Condenser performance characteristics

• Evaporative condenser performance depends on• outside air wet bulb temperature

• as outside air wet bulb temperature increases, evaporative condenser capacity decreases

• capacity decrease is on the order of 2.5% per °F• saturated condensing temperature

• as saturated condensing temperature increases, evaporative condenser capacity increases

• capacity increase is on the order of 6% per °F

Performance characteristics

• Performance factors – cont.• wet/dry operation

• dry operation significantly reduces capacity

• a rule‐of‐thumb is a 65% reduction in capacity in dry vs. wet

• air flow rate• increased air flow rate increases condenser capacity

• increased air flow rate greatly increases condenser fan horsepower


• allow condensing pressure to drop with decreasing outside air wet bulb temperature

• takes advantage of all evaporative condenser capacity during cool outside air conditions

• condensing pressure only allowed to drop to a pre‐determined minimum (for example Pcond,min = 110 psig)


• Consequences of lowering condensing pressure

• increased evaporative condenser energy usage

• decreased compressor energy usage

• reduced high stage compression (on average)

Lowering Condensing Pressure• reduced system energy use

• 0.4‐0.6% reduction for each °F reduction in minimum condensing temperature at Maryland production facility, 0.2‐0.4% reduction for each °F reduction in minimum condensing temperature at Texas facility, 0.5% reduction for each °F reduction in minimum condensing temperature at Florida facility

• increased system capacity

• 0.4% capacity increase for 5 psi condensing pressure reduction (~1.6°F saturation temperature reduction)

• prolonged compressor life & decreased oil cooling loads

• Explore during the winter months

Condensing pressure limits• Limits are dictated by:

• hot gas defrost requirements• setting of defrost relief regulators

• sizing of hot gas main

• condensate management in hot gas main

• DX evaporators• most thermostatic expansion valves need at least 75 psig differential pressure to function properly

• presence of liquid injection oil cooling• check manufacturer’s requirements for TXV pressure differential (limits are relaxed if using motorized expansion valve)

Condensing pressure limits, cont.

• Limits dictated by:

• evaporative condenser selection• close‐approach evaporative condensers usually result in an optimum head pressure that depends on outdoor air temperature (more on this momentarily)

• evaporative condenser fan controls• VFD fans are preferred but 2‐speed fans yield considerable benefits

Condensing pressure limits, cont.

• Limits dictated by:• hand expansion valve settings

• significantly lowering head pressure will likely require seasonal HEV adjustments (liquid makeup to vessels)

• this constraint can be overcome by the use of motorized valves or pulse width valves

• oil separator sizing

• gas driven systems (transfer systems)

• controlled‐pressure receiver set points

• heat recovery

• engineering and operations (knowledge and willingness)

Condensing Pressure Control(Version 1)

• Single speed fan with on/off control• historically most common method of head pressure control

• need to set cut‐in (e.g. 130 psig) & cut‐out pressures (e.g. 125 psig)

• simple control method resulting in• highest energy consumption compared to alternatives

• higher maintenance (fan motors & belts) due to starting/stopping

Condensing Pressure Control(Version 2)

• 2‐Speed fan control• need to set high speed cut‐in (e.g. 135 psig), low‐speed cut‐in pressure (e.g. 130 psig), and low‐speed cut‐out pressure (e.g. 125 psig)

• relatively simple control method resulting in• higher capital cost compared to single‐speed fan option

• lower energy consumption compared to Version 1 control

• sequencing speed controls requires attention

Conclusions

• Lower condensing pressure is GOOD!• If you can’t get your minimum condensing pressure down, you limit your potential savings

• There are limits though…find them for your system!

• Control strategies with VFDs are different that with fixed‐speed fan control

• With VFDs there often is an optimum condensing pressure

• Lower “peak” condensing pressure makes it more pronounced

• The peak occurs at high load & high wet‐bulb temperature

Compressor Efficiency OpportunitiesCompressor Efficiency Opportunities

Reciprocating Compressors

• Compression ratio limits

• 6:1 for splash lubricated wrist pins and castcrankshafts

• 8:1 for rifle drilled connecting rods and shot peened or forged crankshafts

• Systems exceeding these compression ratios require staging

• Today recips are more likely to be seen insmaller or older plants

Rotary Screw Compressors

• Positive displacement

• Single or twin screw

• Compression ratio limits

• capable of 18:1 • practical at 10:1

• One of the fastest growingcompressor types due to size range

Compressor ECMs

• Volume ratio (Vi)

• Oil cooling

• Sequencing & Control strategies• Reduce part‐load operation for screw compressors

Volume Ratio

• Ratio of compressor volume at suction to volume at discharge

• A characteristic of screw compressors

• a given screw compressor may have a fixed volume ratio

• highest pressure in the screw is determined solely by rotor phase & location of discharge port

• If highest pressure is less than the condensing pressure, under‐pressurization occurs

• If highest pressure is greater than the condensing pressure, over‐pressurization occurs

• Both over‐ & under‐compression reduce the efficiency

Vi

Screw compressors are fixed compression ratio devices

Ideally, the Vi will match the compression ratio requirements

k

discharge

suction

suction

discharge

V

V

P

P

VdischargeVsuction

discharge

suction

V

VVi

Example: Fixed Suction Conditions(Tsat,suction = 0°F)

• Given a fixed suction condition (0°F) & a fixed Vi of 3.6:

• any condensing pressures in excess of 160 psig will result in under‐compression

• any condensing pressures below 160 psig will result in over= compression

Pdischarge(psig)

Vi CR

180 3.90 6.40

170 3.75 6.08

160 3.60 5.75

150 3.45 5.42

140 3.29 5.09

130 3.13 4.76

120 2.97 4.43

110 2.81 4.10

100 2.64 3.77

Variable Vi

• Compressor effectively changes location of discharge port to match pressure required by condensing conditions

• Can improve compressor efficiency if fixed Vi is significantly different that required by suction & condensing pressure ratio

• Results in more efficient operation with varying head pressures

• Note that reciprocating compressors are, inherently, variable Vi

Variable Vi

• Vary the compressor’s volume ratio to better match the required compression ratio

k

discharge

suction

suction

discharge

V

V

P

P

Vdischarge

Vsuction

Volume Ratio Control

Volume Ratio Control

Required Volume Ratio

Fixed Volume Ratio Efficiency

Fixed vs. Variable Volume Ratio

Is Variable Vi required for efficient screw compressor operation?

• NO• A well‐chosen fixed Vi screw compressor can perform efficiently over the expected range of condensing pressure

• EXCEPTION: a screw compressor that an swing between multiple suction pressure levels almost assuredly requires Variable Vi for efficient operation at each of the possible suction pressures

• Ok, so what if the system already has Variable Vi?• Look for calibration issues with the control of the Vi• Poor calibration can result in lower efficiency (garbage in…garbage out)

Opportunity

Convert compressors from liquid injection to external oil cooling

Liquid Injection (LIOC) Characteristics

• Evaporation cools refrigerant and oil as it passes through the compressor

• Injects high‐pressure liquid into compressor body

• Liquid feed is controlled to maintain discharge temperature equal to oil supply temperature requirement

• Increased compressor horsepower• Must recompress the evaporated liquid that is injected

• More frequent maintenance on compressor

LIOC Characteristics, cont.

Thermalexpansion

valveHigh-pressureliquid piping

Injection point

External Oil Cooling Characteristics

• Regardless of type, external results in less frequent maintenance on compressor

• Thermosiphon (TSOC)• External heat exchanger required

• Evaporates high‐pressure liquid refrigerant to cool the oil

• Elevated pilot receiver (vessel) usually required• Gravity and buoyancy are the driving forces for liquid feed

• Secondary coolant• Glycol (GOC, cold climates) or water

TSOC

Oil coolingheat exchanger

Supplyliquidpiping

Returnvaporpiping

GOC

Oil coolingheat exchanger

Key Differences

• Energy uses• TSOC – condenser fan+pump energy only• GOC – fluid cooler fan+pump and glycol pump energy• LIOC – compressor + condenser fan energy

• Operational• TSOC & GOC – allows operation at lower head pressures

• Maintenance• TSOC & GOC – less compressor maintenance

• Space• TSOC – more space for oil coolers, elevated pilot vessel, more refrigerant piping

• GOC – space for oil coolers, glycol pump & piping, fluid cooler outside

Typically 2-9% moreCompressor horsepowerfor LIOC

Conservativelyreduced thewinter head pressure setptfrom 135 to120 psig. Nochange in summersetpt.

Case Study Results

• Midwestern Food Processor

• 13 compressors (~5,000 hp)with LIOC

• Two‐stage with 3 suction levels

• ‐45oF, 10oF & 20oF

HSS Loads

MSS Loads

LSS Loads

HPR

CPR

LPA

MPA

HPA

From HPR

From HPR

Case Study Results

• Energy Analysis of conversion resulted in:• ~175 kW peak demand reduction

• 1.1 million kWh (~9%) and $50,000 reduction annually

• Approximately 67% of the savings was energy reduction

• Just under 4 year simple payback on energy costs• without considering maintenance savings and extended compressor life

• received utility rebate for conversion

Other Benefits

• Freed up approximately 100 tons of capacity on the high‐stage suctions from elimination of booster oil cooling load

• LIOC on boosters (i.e. two‐stages of compression) means that the oil cooling load is a high‐stage load

• Oil cooling more available during start‐up• Start‐up with LIOC is more difficult because you have to build up pressure on the high‐side before you get any oil cooling

Oil Cooling Considerations

• consider using pumped glycol rather than thermosiphon• Allows for use of welded plate heat exchangers at the compressor

• No issues siting the elevated thermosiphon pilot receiver

• Easier to balance flows to each compressor oil cooler

• Reduced refrigerant charge required

• Simplified pressure relief protection on oil cooler

• Easy system start‐up because oil cooing is completely independent of refrigeration system pressures

Part‐Load Compressor Performance

• What happens to compressor efficiency when operating at part‐load?

• Reciprocating compressors

• Screw compressors• Single vs. twin

• Fixed vs. variable volume ratio

• How do “system effects” (e.g. pressure drop) alter cataloged performance?

Reciprocating Compressor

0 20 40 60 80 1000

10

20

30

40

50

60

70

80

90

100

Percent of Full Load Capacity

Per

cen

t o

f F

ull

Lo

ad P

ow

er

Recip. Unloading Steps

Ideal Unloading95°F Condensing

compressor‐only

Typical Part‐Load Characteristics

0 10 20 30 40 50 60 70 80 90 1000

10

20

30

40

50

60

70

80

90

100

Percent Capacity

Per

cen

t F

ull

Lo

ad B

HP

Vi=5.0

Vi=3.6

Vi=2.6

(Condensing Temperature > 75 F)

FES Screw Compressor

compressor‐only

Part‐Load Characteristics

0 10 20 30 40 50 60 70 80 90 1000

10

20

30

40

50

60

70

80

90

100

Percent Capacity

Pe

rce

nt

Fu

ll L

oa

d P

ow

er

Vilter Single Screw

compressor‐only

Screw Compressor Part‐Load Operation is Inefficient!

0 10 20 30 40 50 60 70 80 90 1002.0

3.0

4.0

5.0

6.0

7.0

8.0

9.0

Capacity [%]

Eff

icie

ncy

[B

HP

/to

n]

FES 290GL - Variable Vi

-20 F Suction; 90 F Condensing

compressor‐only

Slide valve % does not = Capacity %

0 10 20 30 40 50 60 70 80 90 1000

10

20

30

40

50

60

70

80

90

100

Slide Valve Position [%]

Cap

acit

y P

art

-Lo

ad

[%]

Howden Twin Screw

Variable Speed Screw Compressor

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 110

15

20

25

30

35

40

45

50

55

Part-Load Ratio

Co

mp

ress

or

Po

wer

[kW

E]

Fixed Speed

Variable Speed

FES 315S Booster Compressor

CF Industries

Albany Terminal

Compressor C-2

Fixed Vi=2.6

July 17, 2003

kw=9.55645 + 15.576·PLRCalculated + 26.2308·PLRCalculated2kw=9.55645 + 15.576·PLRCalculated + 26.2308·PLRCalculated2

kw=22.5113 + 11.5401·PLRCalculated + 16.0278·PLRCalculated2kw=22.5113 + 11.5401·PLRCalculated + 16.0278·PLRCalculated2

Linear unloading

Compressor Sequencing

Sequencing Compressor Operation

• Recognize advantages, disadvantages, and limitations of compressor selections

• Make wise choices for fixed Vi screw compressors in high‐stage or single stage systems

• Recips vs. screws?

• Lead screw and lag recip. or lead recip. and lag screw?

• Recognize part‐load characteristics of compressors

Part‐Load Efficiency Comparison

0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.20.2

0.3

0.4

0.5

0.6

0.7

Compressor Part Load Ratio

Co

mp

ress

or

Sp

ecif

ic P

ow

er

Low PressureReceiver

Temperature

-15°F

-5°F

5°F

Reciprocating

Single-Screw

Saturated Discharge Temperatue = 85°F (29.4°C)

(-26.1°C)

(-20.5°C)

(-15°C)

Source: Manske, K. et al., 2000

Efficiency of Two Screw Compressor Operation

Source: Manske, K., et al., 2000

Efficiency of Unequally Sized Screw Compressors

Source: Manske, K., et al., 2000

Compressor Sequencing Basics

• When both screws & recips are available, unload recip firstand screw last

• Always try to operate screw compressors at part‐load ratios greater than 50%

• Note that this may be a slide valve percentage of 60‐70% depending on the compressor

• Operating systems with unequal sized compressors differs from systems with equally sized compressors

Defrosting EvaporatorsDefrosting Evaporators

Frost or no‐frost?

• Frost will form an evaporator when:• The coil surface temperature is below 32°F and,• The entering air dew point temperature is above the coil surface temperature

The Frost Paradox

• Accumulation of frost decreases refrigeration system capacity over time

• Removal of frost decreases refrigeration system capacity during each defrost

• Need to find a compromise between defrost frequency and dwell time

Before You Optimize Defrost

• Eliminate unnecessary sources of moisture• Infiltration of outside air

• Failed seals

• Direct envelope openings

• Plant air imbalance

time

Evap

orat

or c

apac

ityCoil initial condition (no frost)

Coil capacity decreases as frost continues to form

Coil capacity drops rapidly as refrigerant flow is stoppedand the “pump out” process proceeds preparing the coilfor defrost

Parasitic energy is attributed to warming the coil mass and both sensible and latent losses to the space

Hot gas defrost terminates and coil begins to cool down

Coil transitions from a temperature warmer than the space to a temperature cooler than the space so useful refrigeration is now restored

Ideal capacity

Average air velocity ‐ Frosting

0 250 500 750 1000 1250 1500 1750 2000 2250 25001.00

1.25

1.50

1.75

2.00

2.25

2.50

2.75

3.00

3.25

3.50

245

294

343

392

441

490

539

588

637

686

Time [min]

Air

ve

loc

ity

[m

/s]

Run No. 1Run No. 1Run No. 2Run No. 2

Run No. 3Run No. 3

Run No. 4Run No. 4

Air

ve

loc

ity

[f

ee

t/m

in]

Model PredictionModel Prediction

Coil Capacity – Frosting

0 250 500 750 1000 1250 1500 1750 2000 2250 2500

54

72

90

108

126

15

20

25

30

35

40

Time [min]

Co

oli

ng

lo

ad

[K

w]

Co

oli

ng

lo

ad

[to

n]

Run No. 1Run No. 1Run No. 2Run No. 2

Run No.3Run No.3Run No. 4Run No. 4

Model PredictionModel Prediction

Defrost Sequence

Let’s look at typical sequences for defrosting an evaporator

113

ProcessTime[min]

Result System Effect

Pump-out 15Removal of refrigerant from coil in preparation for defrost

Decreasing but positive capacity

Soft-gas 2-10 Slowly raises evaporator pressure Negative load on system

Hot-gas supply

10

Warm coil mass to melt frostNegative load on system (when coil comes out of defrost)

Frost meltNegligible system load – energy leaves system by frost condensate draining

20Excess hot gas beyond what is required to melt frost

Negative load on system while gas continues to be supplied beyond that required to melt frost

Bleed & fan delay 15Pull down coil in preparation for meeting load.

Capacity increases to clean coil capacity over this period

113

Defrost Sequence: Times• Pumpout: 5‐25 minutes

• Soft gas: 5 minutes• Determine by watching pressure in evaporator

• Set for pressure to be 5‐10 psi to the defrost regulator setting of 70 psig

• Hot gas: 15‐30 minutes

• Bleed: 5 minutes• Determine by watching pressure in evaporator

• Set for pressure to be within 5 psi to the suction pressure

• Rechill: 5 minutes

Defrost Sequence: Pumpout

*TRL

DC

T

HG

*TRS

Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series

Goal: evaporate liquid in evaporator so that the pressurewill rise more quickly to defrost

Pumpout Time Estimate

• Consider a Krack 3L‐9610 with 3 fpi operating at ‐30°F evaporating temperature

• Capacity of 4.8 tons per °F TD• 15°F TD gives 72 tons

• Coil Volume of 15.9 ft3

• Assuming the following• 30% full of liquid at beginning of pumpout (±5%)

• 200 lb of liquid ammonia

• 50% of rated capacity during pumpout (±5%)

• Results in an estimated pumpout time for ALLliquid (not practical) of 17±3 minutes

Capacity During Pump‐Out

0 2 4 6 8 10 12 14 16 18 200

5

10

15

20

25

30

Time [min]

coolin

g c

apaci

ty [t

on]

Cooling capacity during pump-out

Defrost Sequence: Soft Gas

*TRL

DC

T

HG

*TRS


Goal: bring the pressure in the evaporator up slowly tolower risk of CIS during in‐rush of HG

Defrost Sequence: Hot Gas

*TRL

DC

T

HG

*TRS


Goal: melt frost from evaporator

Defrost Sequence: Bleed

*TRL

DC

T

HG

*TRS


Goal: slowly reduce pressure in evaporator prior toopening the suction stop valve to suction pressure

Defrost Sequence: Rechill

*TRL

DC

T

HG

*TRS


Goal: freeze any water on evaporator surfaces prior toenergizing fans

Defrost Sequence: Cooling

*TRL

DC

T

HG

*TRS


Goal: cold

Hot Gas Defrost – Energy Flows

1.) Warm mass of coil

2.) Warm mass of accumulated frost to melting point

3.) Change state of frost to liquid

4.) Re‐evaporate portion of liquefied water

5.) Hot gas bypass

Frost melting stages

0 Minute0.5 Minute1 Minute2 Minutes

• cooling mode = 24 hrs

• pump down = 20 min

• Hot gas = 40 min

• bleed = 10 min

• Fan delay = 5 min

2.5 Minutes3 Minutes4 Minutes5 Minutes7 Minutes14 Minutes

Volume flow rate of the melt

0 5 10 15 20 25 30 35 40 450

10

20

30

40

50

60

70

0

2.64

5.28

7.92

10.56

13.2

15.84

18.48

Time min

Vo

lum

e f

low

ra

te [

L/m

in]

Vo

lum

e f

low

ra

te [

ga

l/m

in]

176 Liter

14 Liter(46.5 gal) (3.7 gal)

Down-stream coil average temperature

0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80-40

-35

-30

-25

-20

-15

-10

-5

0

5

10

15

20

25

30

-40

-31

-22

-13

-4

5

14

23

32

41

50

59

68

77

86

Time min

Tem

pera

ture

[oC

]

Run #124hrRun #124hr

Tem

pera

ture

[oF]




Pump-down20 min

Hot Gas 40 min

Bleed10 min

No frost

Convection to Space

• Penthouse Units• Minimizes time required to achieve melt

• Minimizes convective load back to space

• Minimizes re‐evaporation to space• Condensation and frost accumulation will occur on surfaces within the penthouse

• Ceiling Hung• 30‐40% of hot gas supplied can re‐appear as convective/re‐evaporation load

Parasitic Load Due to Excess Hot Gas

• Prolonged supply of hot gas beyond that required for complete defrost will

• Artificially increase load on defrost return suction pressure level

• Increase refrigeration system energy consumption

• Cause suction pressure to cycle – loading and unloading compressors

Optimizing Defrost

• Balances the frequency of defrost• Are multiple defrosts per day needed?

• Seasonally adjust?

• Manage pump‐out

• Manage hot gas dwell period

• Why are you supplying hot gas for more than 15 minutes?

• Do not oversize A4AKs• Seek alternatives to relief regulators

Optimizing Defrost – Ice Cream Storage

0 500 1000 1500 2000 2500 3000-40

-30

-20

-10

0

10

20

30

40

Time [min]

Coil C

apac

ity

[tons]

48 hr cycle

24 hr cycle

During defrost, effective coil capacity is -150 tons

12 hr cycle

Optimizing Defrost

evaporatorQCycle Hot Gas Dwell Capacity evaporator

[hr] [min] [ton-hr] [%]

1210 1506 90

30 1284 76

2410 1465 87

30 1360 81

4810 1284 76

30 1240 74

Conclusions

• Frost accumulates on evaporators operating at low temperatures

• degrades coil performance

• degrades system efficiency

• Critically evaluate your defrost sequences

Variable Frequency Drive ApplicationsVariable Frequency Drive Applications

What are good applications of VFDs in Industrial Refrigeration Systems?

• Condenser FANS? YES. Apply to all fans.

• Condenser pumps? NO!

• Evaporator fans? MAYBE• Dock evaporators? NO

• Storage evaporators? USUALLY

• Blast freezers or spiral freezers? SOMETIMES

• Compressors? MAYBE

Expected Payback Ranges

• Condenser fansoAll or none

o Expect 2‐3% savings

• Evaporator fanso2‐4% savings range

o Simple paybacks 1‐5 years

• CompressorsoAt most, one VFD comp per suction level

o Simple paybacks 1‐4 years

Variable frequency drives

• Good applications• Large motors

• High hours per year operation

• Frequent part‐load operation

• Variable torque processes are best• As speed is reduced, so is torque

• Fans and centrifugal pumps

• Allows application without overheating the motor at low speeds

0

500

1,000

1,500

Ho

urs

per

Year

100 95 90 85 80 75 70 65 60 55 50 45 40 35 30

Percent of Design Load

Good VFD Candidate

Poor VFD Candidate


• Motor requirements• Inverter‐duty may be necessary for variable torque applications (fans)

• Inverter‐duty will be necessary for constant torque applications (compressors)

• VFD requirements & characteristics• Drive must be within 50‐100 ft of application†

• May apply a single drive to more than one motor• Size drive for total connected horsepower• Individual motor over‐current protection required

• Startup torque is reduced

• Power factor• Near unity (1) for VFDs w/harmonics‐mitigating equip.

† manufacturer dependent

VFD Drawbacks

• Drive losses (~2‐5%, losses increase at low loads)

• Additional equipment to maintain

• Resonance of equipment (natural frequency)

• Power quality

• Siting of the drive

Drive Maintenance Considerations

• Clean ‐ keep the drive clean

• Dust and debris reduce air flow through the drive

• Diminished heat removal in the drive will cause premature component failure

• Add a PM to “dust out” your drives – e.g. with compressed air or non‐static sprays

• Dry – keep the drive dry• Moisture and condensation will cause corrosion – particularly on PCBs leading to failure

• Located drive in place that can be maintained dry

Drive Maintenance Considerations

• Connections – keep all connections tight• Connections that become loose due to vibration or thermal cycling can lead to erratic operation and arcing – causing failure

• Create a PM to thermally scan connections (DO NOT RE‐TORQUE CONNECTIONS AS A PM)

• Properly torque connections that are “hot”

• Other• Check with your drive manufacturer for further inspection and maintenance recommendations

VFDs for Refrigeration Compressors

Compressor Capacity Control

• Reciprocating • Start/stop individual compressors (rack system)

• Discrete cylinder unloaders

• Hot gas bypass (not preferred)

• Variable speed drive

• Screw (single & twin)• Continuous slide valve, poppet valves, …

• Hot gas bypass (not preferred)

• Variable speed drive

Capacity Control

Compressor capacity is directly proportional to shaft speed

50 6 0 70 80 90 10050

60

70

80

90

100

Percent C om pressor Speed

Per

cen

t F

ull-

Lo

ad C

ap

acit

y

Twin Screw Compressor

Efficiency Benefit

VFDs perform well at part‐

load conditions!

50 60 70 80 90 1001.5

1.6

1.7

1.8

1.9

2

2.1

2.2

Part-Load Capacity [%]

Eff

icie

ncy

[B

HP

/to

n]

Variable Speed

Fixed Speed (slide valve)

Single Stage15 psi suction

181 psi discharge

thermosiphon oil cooling

Twin Screw

VFD Benefits on Compressors

• Potential for reduced system power

• More efficient compressor performance at part‐load

• More stable suction pressure

VFD Application Considerations

• One VFD‐equipped compressor per suction level in the plant

• Sequence considerations• Lock in fixed speed screws at 100% slide valve

• Trim with VFD‐equipped compressor

• Use speed as first level of capacity control

• Use slide valve as second level of capacity control

• Monitor PI control to avoid speed cycling

• Verify oil circulation system function at low speeds with compressor manufacturer

VFDs for Refrigeration Evaporators

Part‐load evaporator fan operation

• As space load is reduced:• Cycle refrigerant feed, always run fans

• Cycle refrigerant feed, cycle fans after period of time with no call for refrigerant feed

• Raise suction pressure, always run fans

• VFDs

• Which is best?


• Applicable fan laws

• Limitations• Typical minimum motor speeds between 20‐30‐Hz

• Impact on heat exchange

loadfullloadfull CFM

CFM

N

N

3


CFM

hp

hp

76.0


CFM

Capacity

Capacity

Fan horsepower impact of VFD

• Rearranging results in

95.3

95.3

PLRCapacity

Capacity

hp

hp

loadfullloadfull

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

0.4 0.5 0.6 0.7 0.8 0.9 1

PLR

hp

/ h

pfu

ll-l

oa

d

20‐Hz 30‐Hz

VFD benefits on Evaporator Fans

• Reduced system power• Drastically reduced evaporator fan horsepower requirement at part‐load

• Lower refrigeration load from fans• (5‐hp equals 1 ton of refrigeration)

• Potentially fewer system transients

• Increased motor life• Less motor cycling

• Inherently “soft‐start”

VFD benefits (continued)

• Improved power factor (especially on small horsepower motors)

• Decreased noise and “wind‐chill”

• Increased control, more stable temperature control

VFD drawbacks

• Loss of evaporator “throw”

• Typical systems have large number of small evaporator fan motors (cost)

When to considered VFDs

• Load requires close temperature control

• Large fans and motors• Blast freezers, penthouse evaporators with ducting, etc.

• Low TD installations• Not necessarily requiring low TD for space conditions

• Significant & frequently occurring part‐load operation• Northern climates

• High electricity rates

Impact of evaporator liquid feed configuration

• Direct‐expansion• Size thermal expansion valve+distributor and coil circuiting for low load conditions

• Gravity flooded• Good fit because liquid feed is proportional to load

• Overfeed• Liquid supply rate is independent of load

• Suction riser should be sized to overfeed at part‐load conditions

How much can I save?

• Evaporator fan horsepower usually a small fraction of the system horsepower at full‐load

• Low TD load requirements result in larger contribution to the system horsepower & parasitic refrigeration load

• Part‐load• Defined as actual load divided by the installed evaporator capacity

• If no fan control, the fan horsepower contribution to the system horsepower is constant

Fan Speed Control Suction Pressure Control

#1 Fixed Fixed

#2 Fixed Variable

#3 Duty Cycle Fixed

#4 Variable Fixed

#5 Variable Variable

Fan & suction pressure control strategies

Analysis assumptions

• Evaporator• TD = 12oF for cooler and 8oF for freezer

• VFD costs• Assume 5‐hp VFD for each evaporator

• Installation 15 hours/VFD by electrician @ $75/hour

• Energy costs• Blended $0.08/kWh

Compressor + evaporator kW/ton

20 30 40 50 60 70 80 90 1000.6

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

1.5

Percent Load

Co

mp

ress

or

+ E

vap

ora

tor

kW/t

on

Fixed Speed / Fixed SuctionFixed Speed / Fixed Suction

Fixed Speed / Variable SuctionFixed Speed / Variable Suction

Duty Cycling / Fixed SuctionDuty Cycling / Fixed Suction

Variable Speed / Fixed SuctionVariable Speed / Fixed Suction

Variable Speed / Variable SuctionVariable Speed / Variable Suction

Tspace=35 [F]

Nevap=5

TDdesign=11.7 [F]

20 30 40 50 60 70 80 90 1001.6

1.8

2

2.2

2.4

2.6

2.8

3

Percent LoadC

om

pre

sso

r +

Eva

po

rato

r kW

/to

n

Fixed Speed / Fixed SuctionFixed Speed / Fixed Suction

Fixed Speed / Variable SuctionFixed Speed / Variable Suction

Duty Cycling / Fixed SuctionDuty Cycling / Fixed Suction

Variable Speed / Fixed SuctionVariable Speed / Fixed Suction

Variable Speed / Variable SuctionVariable Speed / Variable Suction

Tspace=-20 [F]

Nevap=7

TDdesign=8.5 [F]

Load

profiles

0

500

1,000

1,500

Hours

per

Year

100 95 90 85 80 75 70 65 60 55 50 45 40 35 30


35F cooler

0

500

1,000

1,500

2,000

Ho

urs

per

Yea

r

100 95 90 85 80 75 70 65 60 55 50 45 40 35 30


-20F Freezer

VFD cost

$-

$500

$1,000

$1,500

$2,000

$2,500

$3,000

0.1 1 10 100

Horsepower

VF

D C

os

t p

er

Ho

rse

po

we

r

AF-300 P11, NEMA 1

AF-300 P11, NEMA 4

AF-300 C11

Source: Grainger (Wholesale Price)Manufacturer: Fuji Electric (GE)Pow er supply: 3-phase, 460-VoltApplication: Variable Torque

VFD M ode l

Economic analysis

Cooler (35oF) Freezer (-20oF)

From always on fan control to VFDSavings per ton $75 $120

Capital cost per ton $65† $105*

Installation cost per ton

$55 $80

Simple payback 1.6 years 1.6 years

From cycling fan control to VFDSavings per ton $50 $65

Simple payback 2.4 years 2.8 years

† Purchase of a single 5‐hp VFD to operate all fan motors (2) on evaporator* Purchase of a single 15‐hp VFD to operate on all fan motors (4) on evaporator

Closing thoughts

• Reasonably short payback (<3 years) compared to always running the fan

• Payback can be shorter with evaporators requiring larger horsepower drives

• Longer if cannot use single drive per evaporator• Limit lowest speed to 30‐Hz

• Ask questions prior to implementation• If retrofit

• Is motor compatible with VFD?• Is resonance at lower fan speeds an issue?

• Check actual current draw on motors prior to sizing drive• Fans require and motors can deliver more power at low temperatures

Additional resources

• Northwest Energy Efficiency Alliance Evaporator Fan VFD Initiative• Baseline Market Evaluation Report, April 1999

• Market Progress Evaluation Report No 2., November 2000

• Market Progress Evaluation Report No 2., June 2002

• Reports available at www.nwalliance.org

VFDs for Refrigeration Condensers

VFD benefits on Condensers Fans

• Reduced TOTAL system power

• Potentially fewer system transients

• Increased motor life• Less motor cycling

• Inherently “soft‐start”

Condensing Pressure Control(Version 3.0)

• Variable frequency drive (VFD or VSD, ASD) on fans• need to set a target condensing pressure then fan speed is modulated to maintain set pressure

• ALL condensers fans should be fitted with VFDs & modulated together for maximum benefit

• Block out frequencies that generate fan vibration/failure

• a simple principle and method to implement• higher capital cost alternative

• lower energy consumption than Version 2 control• Fixed target pressure results in many hours at 60‐Hz (i.e. no benefit of VFD)

Condenser fan control map

Strategy Mode 1 Mode 2 Mode 3 Mode 4 Mode 5

Small Motor off on off onLarge Motor off off on onSmall Motor off off onLarge Motor off on onSmall Motor off on on onLarge Motor off off half-speed onSmall Motor off half-speed half-speed on onLarge Motor off off half-speed half-speed onSmall Motor offLarge Motor off

5variable speedvariable speed

1

2

3

4

Comparative cond. fan performance

~44%

~6%

Simple two condenser system

Heat rejection load

Fixed speed control Variable speed drive

# condensers HP#

condensersHP*

100% 2 30 2 @ 100% 30

75% 1 + 1/2 21.6 2 @ 75% 9.8

50% 1 15 2 @ 50% 1.8

Each condenser equipped with 15 HP fan.

* Sans drive losses

Comparative cond. fan performance

~44%

~6%

~32%

~72%


• VFDs on fans• need to specify target wet‐bulb approach, calculate target condensing pressure, and all condenser fan speeds are modulated to maintain set pressure

• more difficult principle and method to implement• highest capital cost alternative

• need to purchase, site, & maintain a wet‐bulb sensor(s)

• harder to determine the target wet‐bulb approach

• lower energy consumption than Version 3.0 control

Version 3.1 Justification

• Version 3.1 was proposed by Manske• based on simulation of a cold storage warehouse with low full‐load, design condensing pressure

• LOTS of condenser capacity

• presented as Master Thesis at UW in 2000

Is there an optimum? Control strategies

Source: Manske, K., 2000

Optimum head pressure control


50 55 60 65 70 75 80120

130

140

150

160

170

180

190

200

210

220

230

1.5x106

1.7x106

1.9x106

2.1x106

2.3x106

2.5x106

2.6x106

2.8x106

3.0x106

3.2x106

3.4x106

Outside Air Wet Bulb Temperature [°F]

Op

tim

um

He

ad

Pre

ss

ure

[p

sia

]

Calculated Ideal Head Pressure (Variable Evaporator Load)Calculated Ideal Head Pressure (Variable Evaporator Load)

Curve Fit (Variable Evaporator Load)Curve Fit (Variable Evaporator Load)

minimum head pressure To

tal S

ys

tem

He

at

Re

jec

tio

n [

Btu

/hr]

Calculated Condenser Heat Rejection (Variable Evaporator Load)Calculated Condenser Heat Rejection (Variable Evaporator Load)

Calculated Ideal Head Pressure (Constant Evaporator Load)Calculated Ideal Head Pressure (Constant Evaporator Load)

Calculated Condenser Heat Rejection (Constant Evaporator Load)Calculated Condenser Heat Rejection (Constant Evaporator Load)

as required by dx txv

Optimum head pressure



• VFDs on fans• need to set target fan speed (usually in 35‐50 Hz range) and all condenser fan speeds are modulated to that speed• set high pressure & low pressure limits and allow modulation of fan speed

away from target speed to maintain those limits

• a simpler principle but still difficult to implement• still a high capital cost alternative (but no wet‐bulb sensor)

• easier to set target speed than approach to wet‐bulb

• harder to switch between speed & pressure control targets

• slightly lower energy consumption than Version 3.1 control

Version 3.2 Justification

• Version 3.2 was proposed by Jekel• based on field evaluation of condenser controls with VFDs & heat recovery

• simulated system with no compressor part load effects

• loosely compared to measured data to verify

• presented at R&T Forum in 2011

Exploring Version 3.2 further

• Consider a 750 ton single‐stage refrigeration system• Three (3) equal sized compressors

• 33.5 psig (20°F saturated) suction pressure

• Single‐speed motors

• Continuous slide‐valve capacity control

• Variable Vi

• Two (2) equal sized evaporative condensers• 25‐hp fan motors with VFD

• 15‐hp water pump motors

• At 78°F design wet‐bulb temperature, system condensing pressure with full‐speed fan operation is 173 psig (92°F saturated)

What does this control look like?

0 5000 10000 15000 2000065

70

75

80

85

90

95

20

30

40

50

60

Condenser Heat Rejection [MBH]

Sat

ura

ted

Co

nd

ensi

ng

Tem

per

atu

re [

°F]

Co

nd

ense

r F

an S

pee

d [

Hz]

WB = 65°FHzset = 45 Hz

SCTSCT

HzHz

SCTmax = 92°F

SCTmin = 70°F

Higher Fan Set Speed

0 5000 10000 15000 2000065

70

75

80

85

90

95

20

30

40

50

60


Sat

ura

ted

Co

nd

ensi

ng

Tem

per

atu

re [

°F]

Co

nd

ense

r F

an S

pee

d [

Hz]


SCTSCT

HzHz

SCTmax = 92°F

SCTmin = 70°F

Lower Fan Set Speed

0 5000 10000 15000 2000065

70

75

80

85

90

95

20

30

40

50

60


Sat

ura

ted

Co

nd

ensi

ng

Tem

per

atu

re [

°F]

Co

nd

ense

r F

an S

pee

d [

Hz]


SCTSCT

HzHz

SCTmax = 92°F

SCTmin = 70°F

Lower Wet‐bulb (45 Hz)

0 5000 10000 15000 2000065

70

75

80

85

90

95

20

30

40

50

60


Sat

ura

ted

Co

nd

ensi

ng

Tem

per

atu

re [

°F]

Co

nd

ense

r F

an S

pee

d [

Hz]


SCTSCT

HzHz

SCTmax = 92°F

SCTmin = 70°F

Full‐load Optimization

20 30 40 50 600.8

0.9

1

1.1

1.2

Evaporative Condenser Fan Speed [Hz]

hp

/to

n o

f C

om

pre

sso

r +

Co

nd

ense

r

WB = 77°FWB = 77°F

WB = 68°FWB = 68°F

WB = 59°FWB = 59°F

WB = 50°FWB = 50°F

Load = 750 tons

Reduced‐load Optimization

20 30 40 50 600.7

0.8

0.9

1

1.1

1.2

1.3


hp

/to

n o

f C

om

pre

sso

r +

Co

nd

ense

r

WB = 77°FWB = 77°F

WB = 68°FWB = 68°F

WB = 59°FWB = 59°F

WB = 50°FWB = 50°F

Load = 550 tons

Reduced‐load Optimization

20 30 40 50 600.8

0.9

1

1.1

1.2

1.3


hp

/to

n o

f C

om

pre

sso

r +

Co

nd

ense

r

WB = 77°FWB = 77°F

WB = 68°FWB = 68°F

WB = 59°FWB = 59°F

WB = 50°FWB = 50°F

Load = 350 tons

Design Weather (hp/ton)

100 200 300 400 500 600 700 8001

1.2

1.4

1.6

Refrigeration Load [tons]

hp

/to

n [

Co

mp

ress

or

+ C

on

den

ser]

WB = 78°F

Hzset = 35 HzHzset = 35 Hz



Design Weather (Fan Speed)

100 200 300 400 500 600 700 80035

40

45

50

55

60


Co

nd

ense

r F

an S

pee

d [

Hz]

WB = 78°F




WB = 63°F (hp/ton)

100 200 300 400 500 600 700 8000.8

0.9

1

1.1

1.2

1.3

1.4


hp

/to

n [

Co

mp

ress

or

+ C

on

den

ser]

WB = 63°F




WB = 63°F (Fan Speed)

100 200 300 400 500 600 700 80010

20

30

40

50

60

70


Co

nd

ense

r F

an S

pee

d [

Hz]

WB = 63°F




WB = 48°F (hp/ton)

100 200 300 400 500 600 700 8000.7

0.8

0.9

1

1.1

1.2

1.3

tons

hp

per

ton



WB = 48°F


WB = 48°F (Fan Speed)

100 200 300 400 500 600 700 80010

20

30

40

50

60

70


Co

nd

ense

r F

an S

pee

d [

Hz] WB = 48°F




Advantages of Version 3.2

• Over Version 3.0 control• Can double the savings of applying VFDs by increasing the number of hours where total system power is reduced (less 60 Hz operation)

• Over Version 3.1 control• No wet‐bulb sensor required (no calibration either!)

• No programming of calculation of target pressure

• Less potential for set point control “hunting”

• Works throughout the year, including partial dry or dry operation

Disadvantages of Version 3.2

• Over Version 3.0 control• More control system programming (true in 3.1 too)

• Over Version 3.1 control• Controlled variable switch

• Between high & low pressure set points, control on fan speed

• Above high pressure set point, control on pressure

• Below low pressure set point, control on pressure

• Stability around the controlled variable switch points

Condenser VFD Conclusions

• Control strategies with VFDs are different that with fixed‐speed fan control

• With VFDs there often is an optimum condensing pressure

• Lower “peak” condensing pressure makes it more pronounced

• The peak occurs at high load & high wet‐bulb temperature

Questions?

Todd Jekel, P.E., Ph.D. Assistant Director, IRC … flooded recirculation Evaporative Condenser(s)...

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Transcript of Todd Jekel, P.E., Ph.D. Assistant Director, IRC … flooded recirculation Evaporative Condenser(s)...