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
HighPressureReceiver
High pressure gas
King valve (automatic)
CondenserEvaporativeCondenser
High pressure liquid
DX Evap 1
Protected suction
…
SuctionTrap
DX Evap n
RefrigerantTransfer System
To HPR
Dis
char
ge
line
CondenserEvaporativeCondenser
MotorizedValve
25 °F
TPController
Power
25 °F
TP
Direct‐eXpansion (DX) System
Evaporativecondensers
Dry suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
Protected suction
…
To high pressure receiverSuction
trap
1
4
Solenoid valve Thermostaticexpansion valve
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
RefrigerantTransfer System
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
Direct‐eXpansion (DX) System
Evaporativecondensers
Dry suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
Protected suction
…
To high pressure receiverSuction
trap
Solenoid valve Thermostaticexpansion valve
Equalizing line
T
DX Evap n
T
DX Evap n
High pressure gasTo plant for evaporatordefrost
RefrigerantTransfer System
Condensers, evaporative
Evaporative condenser, HX
14
Direct‐eXpansion (DX) System
Evaporativecondensers
Dry suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
“Wet” suction
…
To high pressure receiverSuction
trap
Solenoid valve Thermostaticexpansion valve
Equalizing line
T
DX Evap n
T
DX Evap n
High pressure gasTo plant for evaporatordefrost
RefrigerantTransfer System
Receivers, high pressure
Direct‐eXpansion (DX) System
Evaporativecondensers
Un-protected suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
Protected suction
…
To high pressure receiverSuction
trap
Solenoid valve Thermostaticexpansion valve
Equalizing line
T
DX Evap n
T
DX Evap n
High pressure gasTo plant for evaporatordefrost
King valve
RefrigerantTransfer System
King valve
Direct‐eXpansion (DX) System
Evaporativecondensers
Un-protected suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
Protected suction
…
To high pressure receiverSuction
trap
Solenoid valve Thermostaticexpansion valve
Equalizing line
T
DX Evap n
T
DX Evap n
High pressure gasTo plant for evaporatordefrost
RefrigerantTransfer System
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
Direct‐eXpansion (DX) System
Evaporativecondensers
Un-protected suction
Compressor(s)
HighPressureReceiver
TT
High pressure liquid
DX Evap 1
Protected suction
…
To high pressure receiverSuction
trap
Solenoid valve Thermostaticexpansion valve
Equalizing line
T
DX Evap n
T
DX Evap n
High pressure gasTo plant for evaporatordefrost
RefrigerantTransfer System
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
EvaporativeCondenser(s)
Equalizerline
HighPressureReceiver
High pressure liquid
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
HighPressureReceiver
High pressure liquid
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
EvaporativeCondenser(s)
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)
HighPressureReceiver
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)
EvaporativeCondenser(s)
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
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?
Condensing pressure control
• 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
Condensing pressure control
• 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)
Condensing pressure control
• 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
Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series
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
Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series
Goal: melt frost from evaporator
Defrost Sequence: Bleed
*TRL
DC
T
HG
*TRS
Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series
Goal: slowly reduce pressure in evaporator prior toopening the suction stop valve to suction pressure
Defrost Sequence: Rechill
*TRL
DC
T
HG
*TRS
Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series
Goal: freeze any water on evaporator surfaces prior toenergizing fans
Defrost Sequence: Cooling
*TRL
DC
T
HG
*TRS
Bottom‐fed LiquidTop‐fed Hot Gas with Pan in Series
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]
Run #224hrRun #224hr
Run #348hrRun #348hr
Run #448hrRun #448hr
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
Variable frequency drives
• 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?
Variable frequency drives
• Applicable fan laws
• Limitations• Typical minimum motor speeds between 20‐30‐Hz
• Impact on heat exchange
loadfullloadfull CFM
CFM
N
N
3
loadfullloadfull CFM
CFM
hp
hp
76.0
loadfullloadfull CFM
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
Percent of Design Load
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
Percent of Design Load
-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%
Condensing Pressure Control(Version 3.1)
• 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
Source: Manske, K., 2000
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
Source: Manske, K., 2000
Condensing Pressure Control(Version 3.2)
• 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
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 = 55 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
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 = 35 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
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 = 55°FHzset = 45 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
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 = 550 tons
Reduced‐load Optimization
20 30 40 50 600.8
0.9
1
1.1
1.2
1.3
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 = 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
Hzset = 45 HzHzset = 45 Hz
Hzset = 55 HzHzset = 55 Hz
Design Weather (Fan Speed)
100 200 300 400 500 600 700 80035
40
45
50
55
60
Refrigeration Load [tons]
Co
nd
ense
r F
an S
pee
d [
Hz]
WB = 78°F
Hzset = 35 HzHzset = 35 Hz
Hzset = 45 HzHzset = 45 Hz
Hzset = 55 HzHzset = 55 Hz
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
Refrigeration Load [tons]
hp
/to
n [
Co
mp
ress
or
+ C
on
den
ser]
WB = 63°F
Hzset = 35 HzHzset = 35 Hz
Hzset = 45 HzHzset = 45 Hz
Hzset = 55 HzHzset = 55 Hz
WB = 63°F (Fan Speed)
100 200 300 400 500 600 700 80010
20
30
40
50
60
70
Refrigeration Load [tons]
Co
nd
ense
r F
an S
pee
d [
Hz]
WB = 63°F
Hzset = 35 HzHzset = 35 Hz
Hzset = 45 HzHzset = 45 Hz
Hzset = 55 HzHzset = 55 Hz
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
Hzset = 35 HzHzset = 35 Hz
Hzset = 45 HzHzset = 45 Hz
WB = 48°F
Hzset = 55 HzHzset = 55 Hz
WB = 48°F (Fan Speed)
100 200 300 400 500 600 700 80010
20
30
40
50
60
70
Refrigeration Load [tons]
Co
nd
ense
r F
an S
pee
d [
Hz] WB = 48°F
Hzset = 35 HzHzset = 35 Hz
Hzset = 45 HzHzset = 45 Hz
Hzset = 55 HzHzset = 55 Hz
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?