Energy efficient food processing: focus on refrigeration. Refrigeration Systems Review

Transcription

1 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

2 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

3 25 F 25 F Single stage DX system traditional High pressure gas Evaporative Condenser(s) Equalize line 3 King valve 2 High Pressure Receiver 4 High pressure liquid DX Evap 1 DX Evap n T T Un-protected suction 1 Protected suction Equalizing line Suction Trap To HPR Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer System Single stage DX system emerging High pressure gas Evaporative Condenser Equalizer line Evaporative Condenser Discharge line High pressure liquid High Pressure Receiver King valve (automatic) Controller DX Evap 1 P T P T DX Evap n Power Motorized Valve Protected suction Suction Trap To HPR Compressor(s) Refrigerant Transfer System

4 Direct expansion (DX) System High pressure gas Typical conditions: 100 psi < P < 180 psi 130 F < T < 230 F Evaporative condensers 3 Equalizer line King valve High pressure liquid Typical conditions: 100 psi < P < 180 psi 56 F < T < 95 F 2 High Pressure Receiver DX Evap 1 DX Evap n Dry suction 1 T 4 Protected suction T Equalizing line Suction trap To high pressure receiver Solenoid valve Low-side Condition Range: 24 psi < P < 75 psi 10 F < T < 50 F Thermostatic expansion valve Compressor(s) Refrigerant Transfer System Compressor, rotary screw Motor Compressor Motor Oil Separator

5 Oil Separator Motor Compressor Discharge vapor Compressor discharge Oil Oil Separator 1 st Stage Oil Separation 2 nd Stage Oil Separation 9 Compressor, rotary screw Discharge Suction Thermosiphon (refrigerant) oil cooling heat exchanger 10

6 Compressors, reciprocating Discharge Suction Direct expansion (DX) System High pressure gas To plant for evaporator defrost Evaporative condensers High pressure liquid High Pressure Receiver DX Evap 1 DX Evap n T T Dry suction Protected suction Equalizing line Suction trap To high pressure receiver Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer System

7 Condensers, evaporative Evaporative condenser, HX 14

8 Direct expansion (DX) System High pressure gas To plant for evaporator defrost Evaporative condensers High pressure liquid High Pressure Receiver DX Evap 1 DX Evap n T T Dry suction Wet suction Equalizing line Suction trap To high pressure receiver Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer System Receivers, high pressure

9 Direct expansion (DX) System High pressure gas To plant for evaporator defrost Evaporative condensers King valve High pressure liquid High Pressure Receiver DX Evap 1 DX Evap n T T Un-protected suction Protected suction Equalizing line Suction trap To high pressure receiver Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer System King valve

10 Direct expansion (DX) System High pressure gas To plant for evaporator defrost Evaporative condensers High pressure liquid High Pressure Receiver DX Evap 1 DX Evap n T T Un-protected suction Protected suction Equalizing line Suction trap To high pressure receiver Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer 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

11 Evaporator, air cooling Penthouse evaporator in a freezer Ceiling hung evaporator in a dock area DX Fluid Product Silo solenoid TXV external equalize line temperature sensing bulb

12 Direct expansion (DX) System High pressure gas To plant for evaporator defrost Evaporative condensers High pressure liquid High Pressure Receiver DX Evap 1 DX Evap n T T Un-protected suction Protected suction Equalizing line Suction trap To high pressure receiver Solenoid valve Thermostatic expansion valve Compressor(s) Refrigerant Transfer System Transfer System Suction trap Transfer drums Oil pot

13 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

14 Gravity flooded recirculation High pressure gas Evaporative Condenser(s) Equalizer line Solenoid valve Hand expansion valve Flooded evaporator Discharge line 2 3 High Pressure Receiver King valve High pressure liquid Float 1 Protected suction Suction Trap Flooded evap 1 Flooded evap, n Compressor(s) Transfer Station To HPR 4 Gravity flooded air unit evaporator surge drum Fill float Evaporator, vapor return Evaporator, liquid supply HPL Strainer Solenoid HEV

15 Evaporators, liquid chillers Plate and frame liquid chiller Shell and tube liquid chiller Gravity flooded chiller evaporator Plate heat exchanger D C B A

16 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 High pressure gas Discharge line Evaporative Condenser 3 Equalizer line 3 Evaporative Condenser 2 High Pressure Receiver King valve High pressure liquid Wet return Dry suction 1 4 Compressor(s) Pumped recirculator 4 4 T Overfed evaporator(s)

17 Liquid overfeed system Recirculator Float column Pumped liquid line Liquid refrigerant pumps 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

18 Single stage compression, multiple temps Evaporative Condenser(s) Low Temperature Evaporator(s) High pressure receiver Medium Temperature Evaporator(s) Low Temperature Recirculator Medium Temperature Recirculator Low Temperature Compressor(s) High temperature Compressor(s) Two stage compression (single temperature level direct intercooled with single stage liquid expansion) Evaporative Condenser(s) Equalizer High stage discharge gas line 3 King valve Low temperature Evaporator(s) 5 High Pressure Receiver 6 Booster discharge gas line 4 1 Low 5 temperature recirculator 5 Booster suction 7 Intercooler 4 High stage suction 2 Booster Compressor(s) High-Stage Compressor(s)

19 Two stage compression (single temperature level direct intercooled with two stages of liquid expansion) Evaporative Condenser(s) High stage discharge gas line Low Temperature Evaporator(s) High pressure receiver Medium Temperature Evaporator(s) Low Temperature Recirculator Intercooler/ MT Recirc Booster Compressor(s) High-stage Compressor(s) 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

20 Energy Efficiency Low side Opportunities Todd B. Jekel Research Scientist Industrial Refrigeration Consortium Refrigeration efficiency Compressor capacity or horsepower hp/ton hp Capacity, tons Saturated Suction Temperature SST [F] [ F] (i.e. pressure) Compressor hp/ton hp/ton

21 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 onto new suction level Increasing level of difficulty

22 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

23 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

24 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 psig { Water chiller 43 psig 40 F space conditioning 45 psig 35 F cooler 40 psig psig Ice bank 25 psig -25 F freezer

25 ENERGY EFFICIENCY HIGH 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

26 Condensing Pressure ECMs Condensing pressure set point changes Lower condensing pressure set point Add condenser surface area Go to first bullet Increasing level of difficulty Condenser ECMs Fan/pump control Fan VFD (more later) Maintenance Maintain clean surfaces Non condensables (purger)

27 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

28 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

29 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 predetermined minimum (for example P cond,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)

30 Lowering Condensing Pressure reduced system energy use % reduction for each F reduction in minimum condensing temperature at Maryland production facility, % 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)

31 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)

32 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

33 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 Opportunities

34 Reciprocating Compressors Compression ratio limits 6:1 for splash lubricated wrist pins and cast crankshafts 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 in smaller 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 growing compressor types due to size range

35 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, underpressurization occurs If highest pressure is greater than the condensing pressure, over pressurization occurs Both over & under compression reduce the efficiency

36 Vi Screw compressors are fixed compression ratio devices V suction V Vi V suction discharge V discharge P discharge P suction V V suction discharge k Ideally, the Vi will match the compression ratio requirements Example: Fixed Suction Conditions (T sat,suction = 0 F) P discharge (psig) Vi CR 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

37 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 V suction V discharge P discharge P suction V V suction discharge k

38 Volume Ratio Control Volume Ratio Control

39 Required Volume Ratio Fixed Volume Ratio Efficiency

40 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)

41 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

42 LIOC Characteristics, cont. Injection point High-pressure liquid piping Thermal expansion valve 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

43 TSOC Return vapor piping Supply liquid piping Oil cooling heat exchanger GOC Oil cooling heat exchanger

44 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% more Compressor horsepower for LIOC Conservatively reduced the winter head pressure setpt from 135 to 120 psig. No change in summer setpt. Case Study Results Midwestern Food Processor 13 compressors (~5,000 hp) with LIOC Two stage with 3 suction levels 45 o F, 10 o F & 20 o F HPA From HPR MPA HSS Loads MSS Loads HPR From HPR LPA LSS Loads CPR

45 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 highstage 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

46 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?

47 Reciprocating Compressor 100 Percent of Full Load Power compressor only Recip. Unloading Steps 10 Ideal Unloading 95 F Condensing Percent of Full Load Capacity Typical Part Load Characteristics Percent Full Load BHP FES Screw Compressor V i =5.0 V i =3.6 V i = (Condensing Temperature > 75 F) compressor only Percent Capacity

48 Part Load Characteristics Percent Full Load Pow er Vilter Single Screw Percent Capacity compressor only Screw Compressor Part Load Operation is Inefficient! Efficiency [BHP/ton] FES 290GL - Variable Vi -20 F Suction; 90 F Condensing compressor only Capacity [%]

49 Slide valve % does not = Capacity % Capacity Part-Load[%] Slide Valve Position [%] Howden Twin Screw Variable Speed Screw Compressor Compressor Power [kwe] FES 315S Booster Compressor CF Industries Albany Terminal Compressor C-2 Fixed Vi=2.6 July 17, 2003 Fixed Speed Part-Load Ratio Variable Speed Linear unloading kw= PLR Calculated PLR Calculated 2 kw= PLR Calculated PLR Calculated 2

50 Compressor Sequencing Sequencing Compressor Operation Recognize advantages, disadvantages, and limitations of compressor selections Make wise choices for fixed Vi screw compressors in highstage or single stage systems Recips vs. screws? Lead screw and lag recip. or lead recip. and lag screw? Recognize part load characteristics of compressors

51 Part Load Efficiency Comparison 0.7 Compressor Specific Power Single-Screw Reciprocating Saturated Discharge Temperatue = 85 F (29.4 C) Compressor Part Load Ratio Low Pressure Receiver Temperature -15 F (-26.1 C) -5 F (-20.5 C) 5 F (-15 C) Source: Manske, K. et al., 2000 Efficiency of Two Screw Compressor Operation Source: Manske, K., et al., 2000

52 Efficiency of Unequally Sized Screw Compressors Source: Manske, K., et al., 2000 Compressor Sequencing Basics When both screws & recips are available, unload recip first and 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

53 Defrosting 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

54 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

55 Ideal capacity Evaporator capacity Coil initial condition (no frost) Coil capacity decreases as frost continues to form Coil capacity drops rapidly as refrigerant flow is stopped and the pump out process proceeds preparing the coil for defrost Parasitic energy is attributed to warming the coil mass and both sensible and latent losses to the space time 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 Average air velocity Frosting Air velocity [m/s] Time [min] Run No. 1 Run No. 2 Run No. 3 Run No. 4 Model Prediction Air velocity [feet/min]

56 Coil Capacity Frosting Cooling load [Kw] Run No. 1 Run No. 2 Run No.3 Run No. 4 Model Prediction Cooling load [ton] Time [min] Defrost Sequence

57 Process Let s look at typical sequences for defrosting an evaporator Time [min] Result System Effect Pump-out 15 Removal 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 Bleed & fan delay 15 Warm coil mass to melt frost Frost melt Excess hot gas beyond what is required to melt frost Pull down coil in preparation for meeting load. Negative load on system (when coil comes out of defrost) Negligible system load energy leaves system by frost condensate draining Negative load on system while gas continues to be supplied beyond that required to melt frost Capacity increases to clean coil capacity over this period 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: 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

58 Defrost Sequence: Pumpout HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: evaporate liquid in evaporator so that the pressure will 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 ft 3 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 ALL liquid (not practical) of 17±3 minutes

59 Capacity During Pump Out 30 Cooling capacity during pump-out 25 cooling capacity [ton] Time [min] Defrost Sequence: Soft Gas HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: bring the pressure in the evaporator up slowly to lower risk of CIS during in rush of HG

60 Defrost Sequence: Hot Gas HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: melt frost from evaporator Defrost Sequence: Bleed HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: slowly reduce pressure in evaporator prior to opening the suction stop valve to suction pressure

61 Defrost Sequence: Rechill HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: freeze any water on evaporator surfaces prior to energizing fans Defrost Sequence: Cooling HG *TRS Bottom fed Liquid Top fed Hot Gas with Pan in Series DC *TRL T Goal: cold

62 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 cooling mode = 24 hrs pump down = 20 min Hot gas = 40 min bleed = 10 min Fan delay = 5 min 14 Minutes Minutes Minutes 0

63 Volume flow rate of the melt Volume flow rate [L/min] Liter (46.5 gal) 14 Liter(3.7 gal) Volume flow rate [gal/min] Time min Down-stream coil average temperature Temperature [ o C] Run #1 24hr Run #2 24hr Run #3 48hr Run #4 48hr No frost Bleed min Pump-down Hot Gas 40 min min Time min Temperature [ o F]

64 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/reevaporation 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

65 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 Coil Capacity [tons] hr cycle 48 hr cycle 24 hr cycle During defrost, effective coil capacity is -150 tons Time [min]

66 Optimizing Defrost Q evaporator Cycle Hot Gas Dwell Capacity evaporator [hr] [min] [ton-hr] [%] Conclusions Frost accumulates on evaporators operating at low temperatures degrades coil performance degrades system efficiency Critically evaluate your defrost sequences

67 Variable 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

68 Expected Payback Ranges Condenser fans o All or none o Expect 2 3% savings Evaporator fans o 2 4% savings range o Simple paybacks 1 5 years Compressors o At 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 Hours per Year 1,500 1, Percent of Design Load Good VFD Candidate Poor VFD Candidate

69 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 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

70 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 nonstatic 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

71 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

72 Capacity Control Percent Full-Load Capacity Twin Screw Compressor Compressor capacity is directly proportional to shaft speed Percent Compressor Speed Efficiency Benefit Efficiency [BHP/ton] Variable Speed Twin Screw Fixed Speed (slide valve) Single Stage 15 psi suction 181 psi discharge thermosiphon oil cooling VFDs perform well at partload conditions! Part-Load Capacity [% ]

73 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

74 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?

75 Variable frequency drives Applicable fan laws N N full load CFM CFM full load hp hp full load CFM CFM full load Limitations Typical minimum motor speeds between Hz Impact on heat exchange 3 Capacity Capacity full load CFM CFM full load 0.76 Fan horsepower impact of VFD Rearranging results in 1 hp / hp full-load hp hp full load Capacity Capacity full load 3.95 PLR Hz 30 Hz PLR

76 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

77 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

78 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

79 Fan & suction pressure control strategies Fan Speed Control Suction Pressure Control #1 Fixed Fixed #2 Fixed Variable #3 Duty Cycle Fixed #4 Variable Fixed #5 Variable Variable Analysis assumptions Evaporator TD = 12 o F for cooler and 8 o F for freezer VFD costs Assume 5 hp VFD for each evaporator Installation 15 hours/vfd by $75/hour Energy costs Blended $0.08/kWh

80 Compressor + evaporator kw/ton Compressor + Evaporator kw/ton Fixed Speed / Fixed Suction Fixed Speed / Variable Suction Duty Cycling / Fixed Suction Variable Speed / Fixed Suction Variable Speed / Variable Suction T space =35 [F] N evap =5 TD design =11.7 [F] Compressor + Evaporator kw/ton Fixed Speed / Fixed Suction Fixed Speed / Variable Suction Duty Cycling / Fixed Suction Variable Speed / Fixed Suction Variable Speed / Variable Suction T space =-20 [F] N evap =7 TD design =8.5 [F] Percent Load Percent Load 1,500 35F cooler Load profiles Hours per Year Hours per Year 1, ,000 1,500 1, Percent of Design Load -20F Freezer Percent of Design Load

81 VFD cost VFD Cost per Horsepower $3,000 $2,500 $2,000 $1,500 $1,000 $500 $- Source: Grainger (Wholesale Price) Manufacturer: Fuji Electric (GE) Pow er supply: 3-phase, 460-Volt Application: Variable Torque Horsepower VFD Model AF-300 P11, NEMA 1 AF-300 P11, NEMA 4 AF-300 C11 Economic analysis Cooler (35 o F) Freezer (-20 o F) From always on fan control to VFD Savings per ton $75 $120 Capital cost per ton $65 $105* Installation cost per $55 $80 ton Simple payback 1.6 years 1.6 years From cycling fan control to VFD Savings 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

82 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

83 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

84 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 Small Motor Small Motor Small Motor Small Motor Small Motor off off off off off on off on half-speed variable speed off on on half-speed on on on on Large Motor Large Motor Large Motor Large Motor Large Motor off off off off off off on off off variable speed on on half-speed half-speed on on half-speed on

85 Comparative cond. fan performance ~44% ~6% Simple two condenser system Heat rejection load Fixed speed control #condensers HP Variable speed drive # condensers HP * 100% % 30 75% 1 + 1/ % % % 1.8 Each condenser equipped with 15 HP fan. * Sans drive losses

86 Comparative cond. fan performance ~72% ~32% ~44% ~6% 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

87 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

88 Optimum head pressure control Source: Manske, K., 2000 Optimum head pressure Optimum Head Pressure [psia] Curve Fit (Variable Evaporator Load) Calculated Condenser Heat Rejection (Variable Evaporator Load) Calculated Condenser Heat Rejection (Constant Evaporator Load) Calculated Ideal Head Pressure (Variable Evaporator Load) Calculated Ideal Head Pressure (Constant Evaporator Load) minimum head pressure as required by dx txv 3.4x x x x x x x x x x x Outside Air Wet Bulb Temperature [ F] Total System Heat Rejection [Btu/hr] Source: Manske, K., 2000

89 Condensing Pressure Control (Version 3.2) VFDs on fans need to set target fan speed (usually in 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

90 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? Saturated Condensing Temperature [ F] SCT max = 92 F SCT min = 70 F WB = 65 F Hz set = 45 Hz Condenser Heat Rejection [MBH] SCT Hz Condenser Fan Speed [Hz]

91 Higher Fan Set Speed Saturated Condensing Temperature [ F] WB = 65 F Hz set = 55 Hz Condenser Heat Rejection [MBH] SCT max = 92 F SCT min = 70 F SCT Hz Condenser Fan Speed [Hz] Lower Fan Set Speed Saturated Condensing Temperature [ F] WB = 65 F Hz set = 35 Hz Condenser Heat Rejection [MBH] SCT max = 92 F SCT min = 70 F SCT Hz Condenser Fan Speed [Hz]

92 Lower Wet bulb (45 Hz) Saturated Condensing Temperature [ F] SCT max = 92 F SCT min = 70 F WB = 55 F Hz set = 45 Hz Condenser Heat Rejection [MBH] SCT Hz Condenser Fan Speed [Hz] Full load Optimization hp/ton of Compressor + Condenser Load = 750 tons WB = 77 F WB = 68 F WB = 59 F WB = 50 F Evaporative Condenser Fan Speed [Hz]

93 Reduced load Optimization hp/ton of Compressor + Condenser Load = 550 tons WB = 77 F WB = 68 F WB = 59 F WB = 50 F Evaporative Condenser Fan Speed [Hz] Reduced load Optimization hp/ton of Compressor + Condenser Load = 350 tons WB = 77 F WB = 68 F WB = 59 F WB = 50 F Evaporative Condenser Fan Speed [Hz]

94 Design Weather (hp/ton) hp/ton [Compressor + Condenser] Refrigeration Load [tons] WB = 78 F Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz Design Weather (Fan Speed) 60 Condenser Fan Speed [Hz] Refrigeration Load [tons] Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz WB = 78 F

95 WB = 63 F (hp/ton) hp/ton [Compressor + Condenser] Refrigeration Load [tons] WB = 63 F Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz WB = 63 F (Fan Speed) 70 Condenser Fan Speed [Hz] WB = 63 F Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz Refrigeration Load [tons]

96 WB = 48 F (hp/ton) WB = 48 F 1.1 hpperton Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz tons WB = 48 F (Fan Speed) 70 Condenser Fan Speed [Hz] Hz set = 35 Hz Hz set = 45 Hz Hz set = 55 Hz WB = 48 F Refrigeration Load [tons]

97 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

98 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?