Emerging Technologies: VFDs for Condensers. Douglas T. Reindl Director, IRC University of Wisconsin-Madison. University of Wisconsin-Madison

Transcription

1 Emerging Technologies: VFDs for Condensers Douglas T. Reindl Director, IRC University of Wisconsin-Madison University of Wisconsin-Madison 1

2 We ve looked at VFDs on Evaporators and compressors, what is the potential for application on condensers? 2

3 Head Pressure Control Our heat rejection system controls head pressure Evaporative condenser fan controls on/ (single speed fans) two-speed fans variable speed fans 3

4 Floating Head Pressure Control Consequences of lowering head pressure increased evaporative condenser energy use decreased compressor energy use reduced high stage compression (on average) Does the decrease in compressor energy use outweigh the increase in condenser fan energy use??? 4

5 Optimum Head Pressure Axial Fan T Compressor+Condenser oa,wb =78 F Power (kw) Compressor T cond,opt = 87.1 F Condenser Saturated Condensing Temperature (F) 5

6 Optimum Head Pressure 300 Centrifugal Fan T oa,wb =78 F 250 Power (kw) Compressor+Condenser Compressor Condenser T cond,opt = 89.9 F Saturated Condensing Temperature (F) 6

7 Optimum Head Pressure Depends on: Condenser fan type (axial vs. centrifugal) Fan control strategy Condenser sizing strategy 95 F saturated condensing (historic) 90 F saturated condensing (recommended) 85 F saturated condensing (possible not practical) 7

8 Case Study: Cold Storage Warehouse Size 34 F 39,000 (ft²) 0 F 52,000 (ft²) 600,000 (lbs/day, food) Type ammonia, single-stage compression, liquid overfeed evaporators Operating Costs 9,000 ($/month) 4 Compressors Available Instrumentation Temp, Pressure, Mass Flow! Defrost Strategies Head Pressure Control The refrigeration system examined as part of this case study is a cold storage warehouse facility located near Milwaukee, WI. The facility contains four types of refrigerated spaces low temperature freezer (0 F), cooler (34 F), docks (45 F), and ripening rooms (45-64 F). From a thermal mass perspective, the warehouse construction type can be considered lightweight for all spaces. There is mostly insulation and very little mass in the walls and roofs. The freezer and cooler with its loading dock are separate buildings located adjacent to each other. The banana and tomato ripening rooms are located in a heated space adjacent to the cooler. The refrigerant used throughout this system is ammonia (R-717). Evaporators in the freezer are top fed, pumped liquid overfeed. Cooler, and cooler dock evaporators are all bottom feed pumped liquid overfeed where as the evaporators in the banana and tomato ripening rooms are direct expansion controlled by thermal expansion valves and back pressure regulators. (NOTE: This case study was conducted by Manske, K. A. in partial fulfillment of the requirements for a MS degree in Mechanical Engineering under the direction of Professor s Reindl, D. T., and Klein, S.A. during Portions of the thesis prepared by Mankse titled Performance Optimization of Industrial Refrigeration Systems, 1999 have been excerpted for this section. A complete copy of the Manske thesis is available for download at: 8

9 Case Study: Cold Storage Warehouse Q reject HPR Condenser F D X Evap Q spac e Design Loads BPR Fruit Ripening = 90 tons Cooler = 107 tons Freezer = 106 tons PLO Evap Q spac e PLO Evap Q spac e 23 F -10 F Yearly Average Loads Fruit Ripening = 43 tons Cooler = 58 tons Freezer = 71 tons There are three main vessels in the system as shown above. The first is the high pressure receiver where liquid refrigerant draining from the condenser is stored. Liquid refrigerant from the high pressure receiver is then throttled to either the intermediate pressure receiver or to the direct expansion evaporators in the banana and tomato ripening rooms. The temperature of the refrigerant in the banana/tomato room evaporators is regulated at a desired level by use of a back-pressure regulator. The back-pressure regulator then throttles the refrigerant gas to the intermediate pressure receiver which is at a lower temperature/pressure. Liquid in the intermediate pressure receiver is then either pumped to the cooler and cooler dock evaporators or throttled again to the low pressure receiver. Liquid refrigerant from the low pressure receiver is pumped to freezer evaporators with a mechanical liquid recirculating pump. Liquid levels in the intermediate and low pressure receivers are maintained at a near constant level by a pilot operated, modulating expansion valve controlled by a float switch located on the receiver tank. A single-screw (Vilter model# VSS 451 connected to the low temperature vessel) and reciprocating compressor (Vilter model# VMC 4412 connected to the high temperature vessel) operate in parallel, each compressing to a common discharge header and a single evaporative condenser. The suction line from the low pressure receiver leads to the screw compressor. The suction line from the intermediate pressure receiver leads to the reciprocating compressor. Additional compressors, in parallel piping arrangements to the primary compressors, can be brought on-line if the load exceeds the capacity of the primary compressors. 9

10 Control Options Single speed fan with on/ control most common head pressure control method set cut-in (e.g. 150 psig) & cut-out pressures (e.g. 140 psig) simple control method but results in higher energy consumption vs. two-speed or VFD higher maintenance (fan motors & belts) liquid management problems w/multiple condensers 10

11 Control Options Cont. 2-Speed fan control set high speed cut-in (e.g. 160 psig) low-speed cut-in pressure (e.g. 150 psig), and low-speed cut-out pressure (e.g. 140 psig) relatively simple control method but results in higher capital cost compared to single speed fan option lower energy consumption vs. single-speed but slightly higher energy consumption compared to variable speed yields less system transients compared to single speed sequencing speed controls requires attention 11

12 Control Options Cont. Variable speed fan (VFD) set a target head pressure modulate fan speed to maintain head pressure a very simple principle & method to implement highest capital cost alternative lowest energy consumption control alternative modulate all condensers the same in systems with multiple evaporative condensers results in smoother system operation with minimal transients 12

13 Condenser Fan Control Map Strategy Mode 1 Mode 2 Mode 3 Mode 4 Mode Small Motor half-speed half-speed on on Large Motor half-speed half-speed on Small Motor Small Motor Small Motor Small Motor on on variable speed on on on on Large Motor Large Motor Large Motor Large Motor on variable speed on on half-speed on on The above map provides five different strategies that could be used for an evaporative condenser that is equipped with twin motors, two-speed fans, or variable speed fans. The modes are indicative of changes in head pressure (either increasing as one moves from left to right or decreasing as one moves from right to left). For example, strategy 3 would work as follows. In mode 1 all fans are. As the head pressure rises, the system responds by energizing a small fan motor in attempts to maintain system head pressure. If the head pressure continues to rise and the setpoint is not satisfied, mode three is initiated by the start of the larger fan motor to half-speed. As the head pressure rises further, mode 4 dictates that the larger fan motor is tripped to run at high speed. The exact opposite sequence occurs as the head pressure falls. 13

14 Condenser Fan Control Options The above figure illustrates the required fan energy (expressed as a percentage of full-load fan power) as a function of the evaporative condenser capacity for the five strategies listed previously. The least efficient option is the on/ control (strategy 1) while the most efficient option is the variable speed drive option. The two-speed fan option yields nearly all of the part-load power and capacity benefits of the variable speed option but with much less costly equipment. Notice that at zero fan power for all options, the capacity of the evaporative condenser is not zero. This is due to the fact that natural convection will occur drawing air through the condenser coils and rejecting heat yielding about 10% of the condenser s heat rejection capacity while the fans are idle. This assumes that the condenser coils are running wet i.e. water continues to flow over the condenser coils. 14

15 Condenser Fan Controls May Source: Manske, K., 2000 Of course we do not want to just minimize the power of the evaporative condenser at the expense of the system; consequently, we must look at the impacts or trades associated with spending more energy on evaporative condenser fans vs. the reduction in compressor power that accrues due to the lower head pressure. The case study system had an oversized evaporative condenser. As a result, it was possible to drive head pressures extremely low in the system. So low in fact that the incremental expenditure of fan energy was not compensated for by an incremental reduction in compressor energy demand. The above plot shows the comparison between heat rejection system control strategies. The point furthest to the left on the curves in the figure represents the system balance point head pressure at which the condenser is operating at 100 percent capacity (for a given outdoor air wet bulb and system load during a peak hour on a average day in May). Any further decrease in condensing pressure would prevent the condenser from rejecting the required amount of energy from the system. The figure shows that VFD fan control could save the system nearly 8% in combined compressor and condenser energy requirements if the head pressure were raised to 125 psia. VFD fan control looses its advantages at low head pressures because the fans must run at near full speed most of the time anyway. At high head pressures the fans in on/ control don t stay on long because of the high rate of heat transfer that occurs. However, at high head pressure an on/ control strategy would cycle the fans on and frequently which would cause excessive wear on the motors and fan belts. The figure also shows that there is a different optimum head pressure for each type of condenser fan control. It is also interesting to note that half-speed fan motors have energy requirements that are only approximately one percent above the VFD motors at elevated head pressures. Since this system has a minimum allowed head pressure of 130 psia, VFD and half-speed motors may have very similar energy requirements for most of the year. 15

16 Optimum Head Pressure Control Source: Manske, K., 2000 This plot illustrates the preferred control head pressure control strategy for two different evaporative condenser sizes. With an evaporative condenser sized for 95 F saturated condensing temperature on a design day, the optimum head pressure is the lowest head pressure achievable by running the evap condenser fans full out. If the condenser is oversized (i.e. an oversized evap condenser is defined as one that yields a saturated condensing temperature of 85 F on the design day), there is an optimum head pressure (i.e. a head pressure greater than the minimum achievable that will minimize the combined power of the compressor and condenser). In this case, the optimum head pressure is likely a function of the outside air wetbulb temperature. The dark set of lines is for the condenser that is currently installed in the system. The current condenser is large enough to allow the system to balance out with a saturated condensing temperature of 85 F on the design day. The compressor/condenser power with a smaller condenser is given by the lighter colored line. The point furthest to the left on each line represents the pressure at which the evaporative condenser has reached 100 percent capacity. Given that the load is constant, it would be physically impossible to achieve a lower head pressure without adding additional condensing capacity. Note, the above case assumes that the refrigeration load is progressively decreasing during the winter months; however, refrigeration load has little influence on the optimum head pressure. Because of the presence of high temperature direct-expansion coils in the case study system, the head pressure is not allowed to go below 130 psia. Therefore, the system cannot possibly be operated at its ideal head pressure except for the months of June through September. It must be operated above its optimum head pressure resulting in a slight excess of compressor power. 16

17 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 When performing the calculations to identify the optimum condensing pressure for the year, we discovered that the optimum condensing pressure had a near linear relationship with the outside air wet bulb temperature. The above curve illustrates the relationship between optimum head pressure and outside air wetbulb temperature (lower curve) over a range of evaporator load conditions (corresponding variability in heat rejection is shown by the points above). In the case of this system, a very simple linear relationship was developed that allows a supervisory reset on the system head pressure given the prevailing outside air wet bulb temperature according to the following: P head,opt = * T wb where P head,opt is the head pressure corresponding to minimum system power in psia and T wb is the outside air wet bulb temperature in F. This relationship assumes that the condensers have variable speed drives. Keep in mind that the above relationship needs to have a lower bound as dictated by the characteristics of each given system. 17

18 Optimizing Head Pressure 1. Measure the outdoor air wet bulb temperature 2. Note the current condensing pressure and system electrical demand 3. Reset the condensing pressure down 5 psig & allow system to equilibrate 4. Note the new system electrical demand 5. Continue steps 3 and 4 until the lower condensing pressure limit setpoint is reached 6. Plot the system electrical demand vs. the condensing pressure and note the condensing pressure corresponding to point of minimum system electrical demand 7. Plot that single optimum condensing pressure point on a optimum condensing pressure vs. outdoor air wet bulb temperature curve 8. Repeat the procedure from 1-7 to more fully develop a curve analogous to the figure given on the previous page. Procedure for Determining Optimum Relation Between Condensing Pressure and Outdoor Wetbulb The trajectory of optimum condensing pressures for corresponding outside air wet bulb temperatures as shown on the previous page is specific to the existing ammonia system. Each system will have its own unique trajectory. However, the following procedure can be used to empirically develop the trajectory of optimum condensing pressures. Note, this procedure needs to be executed during -design periods of the year (during relatively lower outside air wet bulb conditions). The procedure also requires the ability to continuously monitor the outdoor air wet bulb temperature, condensing pressure, and the engine room total electrical demand. We also recommend that other system state variables (such as suction pressures, superheat if applicable, etc.) be monitored to ensure reliable system operation during the procedure. 1. Measure the outdoor air wet bulb temperature 2. Note the current condensing pressure and system electrical demand 3. Reset the condensing pressure down 5 psig (35 kpa) and allow the system to equilibrate 4. Note the new system electrical demand 5. Continue steps 3 and 4 until the lower limit in condensing pressure setpoint is reached 6. Plot the system electrical demand vs. the condensing pressure and note the condensing pressure that corresponds to the point of minimum system electrical demand 7. Plot that single optimum condensing pressure point on a optimum condensing pressure vs. outdoor air wet bulb temperature curve 8. Repeat the procedure from 1-7 to more fully develop a curve analogous to the figure given on the previous page. Once the optimum condensing pressure trajectory curve is developed, it can be programmed i t t PLC i t ll t i ld ti t f th h t 18

19 Economic Benefits of Drive Annual savings for warehouse 13 kw (peak) reduction 97,140 kwh reduction (~5%) $3,856 per year in electrical operating costs (~5%) Drive cost = $6,900 Simple payback of 1.8 yrs 19

20 Final Thoughts VFDs on condensers can provide economic and operating cost benefits on the high-side Take advantage of lowering head pressure Consider barriers to lowering head pressure 20

21 Floating Head Pressure Control Head pressure limits dictated by: hot gas defrost requirements setting of defrost relief valves 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 liquid injection oil cooling check manufacturer s requirements for TXV pressure differential As with most things, there are limits to lowering system head pressure. We do not want to create problems by trying to improve the efficiency of our systems. The above items are some of the more common factors constraining or limiting our ability to lower system head pressure. Keep in mind that these items may not necessarily be unmovable barriers; however, changes in components or system arrangements may be required to overcome their limiting effects on the system. Hot Gas Defrost: Many industrial refrigeration systems utilize hot gaseous refrigerant to defrost evaporators. In cases where defrost relief valves are installed, a sufficient pressure differential (e.g. 75 psig) across the valve must be created to open the valve. Sizing of the hot gas main may also impose constraints. If a hot gas main is undersized, hot gas (at a sufficient rate) will not be delivered to the evaporator without a high differential pressure. For larger size hot gas mains, a much lower differential pressure will allow adequate flow of hot gas to defrosting evaporator(s). Finally, if condensate is not properly managed in hot gas mains, hydraulic shock can cause catastrophic failures of hot gas piping on a call for defrost. Also, the condensate effectively decreases the pipe size causing similar symptoms as an undersized line with regard to head pressure requirements. All of these deficiencies can be overcome in the long run; however, they do create real barriers to lowering head pressure in the short run. DX Evaporators: In systems that utilize direct-expansion evaporators, a minimum differential pressure is required across the thermostatic expansion valves (TXV). The minimum pressure differential is dependent on the specific valve selection but is routinely on the order of 75 psig. When we drop the pressure differential across the TXV below the minimum, we lose controllability of the valve (control engineers call this control authority ). What results is an inability to properly modulate refrigerant to the evaporator. Since our evaporator pressure i.e. downstream of the TXV the pressure is fixed (to satisfy our temperature requirements for meeting load), the head 21

22 Floating Head Pressure Control Head pressure limits dictated by: evaporative condenser selection oversized evaporative condensers results in an optimum head pressure that depends on outdoor air wet bulb temperature evaporative condenser fan controls VFD fans are preferred but 2-speed fans yield considerable benefits thermosiphon oil coolers Evaporative Condenser Selection: If an evaporative condenser is too small, the system head pressure will rise until its heat rejection capacity is sufficient to reject the needed heat from the system. Fan Controls: Although fan controls themselves do not necessarily limit head pressure, there are methods of fan controls that lead to more stable and efficient system operation. Two speed condenser fans or variable frequency drive (VFD) fans have better capacity modulating capability and result in more stable head pressures leading to more stable system operation. In addition to their stability, two-speed and VFD controlled fans will result in improved energy due to their better part-load performance as compared to single speed fans. Thermosiphon Oil Cooling (TSOC): TSOC improves compressor efficiency by using a thermosiphon effect coupled with the system s evaporative condenser to reject heat from the compressor s oil. 22

23 Floating Head Pressure Control Head pressure limits dictated by: hand expansion valve settings significantly lowering head pressure will likely require seasonal HEV adjustments this constraint can be overcome by the use of motorized valves or pulse width valves oil separator sizing gas driven systems (transfer systems & gas pumpers) controlled-pressure receiver setpoints heat recovery engineering and operations (knowledge & willingness) 23