By Eric M. Moe
About the Author
Eric M. Moe is director of business development at Flow Control Industries in Woodinville, Wash.
Building Performance
With District Cooling
L
ow chilled water temperature differential (
∆T) continues to
T) continues to
T
be a major district cooling weakness. Poor
L
be a major district cooling weakness. Poor
L
∆T performance
T performance
T
at cooling coils leads to lost cooling capacity, wasted energy, extra
cost and added complexity for a thermal utility, its chilled water
customers, or both.
To encourage customers to invest in technology to improve ∆T performance T performance T in their buildings, an increasing number of utilities have established chilled water rates that vary inversely with ∆T. T. T Figure 1 is an example of rates charged to custom-ers from one prominent univcustom-ersity in the United States. As you can see, the lower the ∆T, the greater the rate. Conversely, T, the greater the rate. Conversely, T customers that minimize their fl ow rate per ton cooling are rewarded.
Customers want to keep their occupants comfortable, minimize total energy cost, and reduce the equipment required for district cooling service. Unfortunately,
some efforts to reduce the fl ow simply shift issues from the central plant to the building. If the approach to raising ∆T reduces cooling coil capacity, increases building pump and fan energy consump-tion, or adds complexity, the benefi t of lower chilled water rates for chilled water customers is marginal at best.
Rapid district cooling industry growth depends on the ability to satisfy existing customers and effectively separate the advantages of district energy from self-generated heating and cooling. Choosing an approach that optimizes performance at cooling coils can simplify building
op-eration, reduce capital expenses, increase system reliability, improve diversity, and consume less (total) energy.
In a large chilled water system, the root cause of low ∆T is typically within your T is typically within your T customer’s building at the load. Cooling coil and control valve performance must be considered to fully address system-wide performance issues.
Chilled Water Customer Interconnections
strategies may return water to the plant at less than design
strategies may return water to the plant at less than design
temperature.
To achieve excellent performance in the building, distribu-tion, and central plant, it is critical for cooling coils to generate high ∆T through the range of load conditions. Furthermore, T through the range of load conditions. Furthermore, T chilled water supply temperature (CHWST) shouldn’t rise to a level that compromises coil capacity, causes humidity problems, or creates a net rise in the total cooling energy consumed.
Direct connected building with booster pump: adds head
to the supply to ensure fl ow through coils; it may take fl ow from or create reverse fl ow at other buildings.
Direct connected building with pressure knockdown valve: reduces the inlet pressure to coils and control valves.
Decoupled building with both supply temperature and differential pressure control: uses the valve in the chilled
water supply line to regulate the inlet pressure and the valve in the crossover bridge to regulate an increase in the CHWST.
Decoupled building with return water temperature con-trol: circulates chilled water in the building until the return
water temperature rises high enough to return it to the plant, which is intended to compensate for poor ∆T performance T performance T across coils.
Decoupled building with heat exchanger: completely
separates the building from the distribution fl ow; it creates additional heat transfer loss.
Direct connected building: simplest approach requiring
the least amount of equipment and physical space in the
build-ing; it is susceptible to changes in pressure with conventional ing; it is susceptible to changes in pressure with conventional control valves.
Direct connected building with booster pump and bypass:
simple approach boosts the supply pressure but only if required at high load; it is susceptible to changes in differential pressure with conventional control valves.
Decoupled Building with Return Water Temperature Control Decoupled Building with Return Water Temperature Control
Return water temperature control is a common strategy used in buildings served by large district cooling systems. By decoupling the building fl ow from the primary or secondary distribution fl ow, chilled water is recirculated in buildings until enough heat has been gained to produce adequate ∆T (T (T Figure (Figure ( 2 [Scheme D] and Figure 3).
While this technique has been used as an intermediate step to generate the desired temperature rise at the plant or through the main distribution pipes, it is fraught with potential problems for customers. When return water is permitted to blend with supply, CHWST increases. This commonly leads to building performance issues including wasted pump and fan energy, lost coil capacity, and poor humidity control. On a system level, pumps used in buildings close to the plant may starve buildings downstream of the pressure and fl ow required to serve the load.
Dirty fi ns and coils, high chilled water supply temperature, low supply air temperature, improper piping, over-pumping, poorly modulating control valves, improper valve sizing, and control valve hunting are a few signifi cant factors that will adversely affect chilled water ∆T performance. T performance. T
A signifi cant body of published literature discusses low ∆T problems as well as design, operation, maintenance and control strategies typically used to attempt to mitigate the issues. De-spite the industry’s best efforts, low ∆T continues to be an un-T continues to be an un-T resolved problem in the majority of large facilities served with central cooling equipment. Although control valve selection and performance is a key contributor to total system performance in operation, it seldom receives the attention it deserves. Chilled Water Supply Temperature
Chilled Water Supply Temperature
Figure 4 illustrates how ∆T, fl ow, and capacity of a 45/55 T, fl ow, and capacity of a 45/55 T split ARI-certifi ed cooling coil varies with a change in the chilled water supply temperature. At 100% design fl ow for the coil, observe the dramatic reduction in capacity when 0.50 0.40 0.30 0.20 0.10 0.00 R
ate ($ per ton/h)
0 5 10 15 20 25
Chilled Water ∆T (°F)T (°F)T
gpm = 24 ton ∆T
Figure 1: Incentive chilled water rates.
A B C D E F G
Figure 2: Common chilled water interconnections to buildings.
140% 120% 100% 80% 60% 40% 20% 0% 0% 20% 40% 60% 80% 100% 120% % Design Flow CHWST 39°F 45°F 42°F 48°F 20°F, 15°F, 10°F ∆T % T otal Capacity 90 psi 90 psi 30 psi 30 psi Distribution Head 40°F 54°F 45°F Pump Suction Return Water Blends With Supply Primary or Secondary Pumps 54°F Supply Water Blends With Return 54°F Return Water Temp. Control Valves No Flow Through Bridge 105 psi Bldg. A Bldg. B Bldg. C 15 psid 20 psid 50 psi
Figure 3 (left): Return water temperature control. Figure 4 (right): Effect of rising CHWST on coil capacity, fl ow rate and ∆T. 40°F
Distance from Central Plant the central plant plotted along the horizontal axis. Pumps develop head. Chilled water fl ows from high to low pressure. In the hy-draulic gradient shown in Figure 3, plenty of distribution head exists to serve the load without running the building pumps close to the plant.2 The following describes how each building in the system is operating.
• Building A: a portion of the chilled water return blends with supply in an effort to reach a 54°F (12.2°C) return water temperature set-point. CHWST to the cooling coils increases. • Building B: a portion of the supply water bypasses coils in the building when the return water temperature leaving the coils exceeds 54°F (12.2°C).
• Building C: no chilled water fl ows through the crossover bridge.
It is easy to see from Figure 3 that the return water temperature control valves for the buildings consume all the remaining head developed by the building pumps, as well as
any additional pressure between the supply and return header. For the load condition shown, none of the pumps in Buildings A, B and C must be run.
When all building pumps are run, as shown in Figure 3, the suction head pressure differential in the distribution steps down as it passes each additional building on the loop. This process can tend to starve buildings downstream of the pressure dif-ferential and fl ow that they need. Larger building pumps that generate more suction head are not the solution to generate chilled water fl ow in starved buildings.
the CHWST rises to 48°F (8.9°C). Also, observe the high ∆T that the coil is capable of producing when low CHWST T that the coil is capable of producing when low CHWST T is maintained.
Proper Control Valve Sizing/Rules of Thumb Proper Control Valve Sizing/Rules of Thumb
The hydraulic gradient in Figure 5 is used to illustrate a simple variable primary fl ow system with direct return piping. As shown, the closer a circuit
is to the plant, the greater the differential pressure that is consumed by the coil and con-trol valve between the supply and return header.
ASHRAE is reevaluating recommendations regard-ing control valve selection.3 None of the current recom-mendations take into account dynamic system performance
or consider future variations in the hydraulic profi le as sys-tems expand. Control valves are commonly sized for peak
load conditions using rules of thumb without taking into consideration the differential pressure at each circuit. Some of the most common methods for control valve selection this author has seen used are:
• Size all control valves for 5 psi (~12 ft [~36 kPa]) pres-sure drop;
• Select control valves for the same pressure drop as the coil it serves;
• Select control valves for line size or one smaller;
• Don’t choose control valves smaller than half the line size;
• Choose control valves for 25% – 50% of the system pressure; and
• Ensure the valve authority is relative to the pressure drop in the circuit served.
Using Figure 5, consider the impact of using the same pres-sure drop to size each coil control valve when design fl ow for
Circuit Near Plant Most Remote Circuit
Total Pressure Drop 96 ft 20 ft Coil Pressure Drop 8 ft 8 ft Valve Pressure Drop 88 ft 12 ft Coil Design Flow 225 gpm 225 gpm Wide Open Valve 617.2 gpm 227.9 gpm Percent Flow
Through Valve 36.9% 100% At Coil Design Flow
Table 1: Impact of sizing all control valves for 12 ft pressure drop.
Total Load ∆T T T CHW Flow CHW Flow Dist. Drop Dist. Drop Coil Drop Coil Drop Valve Drop Valve Drop Total PD Total PD Pump Power Pump Power Motor PowerMotor Power
350 tons 18.0°F 467 gpm 1.22 ft 1.39 ft 12.00 ft 14.6 ft 2.0 bhp 1.7 kW 350 tons 15.0°F 560 gpm 1.75 ft 2.00 ft 12.00 ft 15.8 ft 2.6 bhp 2.2 kW 350 tons 12.0°F 700 gpm 2.73 ft 3.13 ft 12.00 ft 17.9 ft 3.7 bhp 3.1 kW 350 tons 7.5°F 1120 gpm 7.00 ft 8.00 ft 12.00 ft 27.0 ft 9.0 bhp 7.5 kW
Sensible Load Exhaust Air Supply Air Airfl ow Dist. Drop SP Setpoint Total PD Fan Power Motor Power
280 tons 75°F 55.0°F 154,839 cfm 2.08 in. H2O 1.00 in. H2O 3.1 in. H2O 100.7 bhp 83.5 kW 280 tons 75°F 60.0°F 206,452 cfm 3.70 in. H2O 1.00 in. H2O 4.7 in. H2O 203.6 bhp 168.7 kW 280 tons 75°F 62.0°F 238,213 cfm 4.93 in. H2O 1.00 in. H2O 5.9 in. H2O 294.8 bhp 244.4 kW 280 tons 75°F 62.1°F 240,000 cfm 5.00 in. H2O 1.00 in. H2O 6.0 in. H2O 302.1 bhp 250.4 kW
Table 2: Building pump motor power requirements with decreasing ∆T.
Table 3: Fan cfm and motor power requirements with rising SAT. Pump Discharge Loss
Chiller Loss Pump Head
Closest Circuits
Furthest Circuit
Pump Suction Loss 190 170 150 130 110 90 70 50 Distance Distance System P ressure (ft)
One Week (Summer 2003) One Week (Summer 2003) Building 1 Building 2 40 35 30 25 20 15 10 5 0 Building Differential P ressure (psid)
Figure 8: Pressure independent modulating two-way control valve. Control Shaft
Rotates to Modulate Flow Rotates to Modulate Flow
Springs Springs Springs Piston Flow Flow P33 P1 P1 Flow Flow Seal Seal Control Surfaces Control Surfaces Flow Flow P2
each coil is 225 gpm. It is easy to see how control valves close to the plant can become grossly oversized for the application. In Table 1, control valves were selected with a Cv of 100 to deliver at least 225 gpm (14.2 L/s) design fl ow assuming a 12 ft (36 kPa) pressure drop across each control valve. No consideration was given to the location of the valve in the system.
At design fl ow through the coil, the control valve in the circuit near the plant is only fl owing 36.9% of its capabil-ity. If this is a commercial quality control valve with typical construction and range, ∆T performance will suffer, especially T performance will suffer, especially T at part load.
Uncertainty in the Hydraulic Profi le Uncertainty in the Hydraulic Profi le
It often is not well understood what differential pressure to use to size conventional control valves. In many cases, the ab-sence of this information leads to control valve sizing by rules of thumb. Figure 6 illustrates differential pressure data from Figure 6 illustrates differential pressure data from Figure 6 two buildings on a university campus. Knowing the system has low ∆T, and the system is expanding, what is the best strategy T, and the system is expanding, what is the best strategy T to size control valves for these buildings? Even if conventional control valves are properly selected now, the hydraulic profi le may change in the future.
Control Valve Hunting
If, at a specifi c cooling coil, the differential pressure changes across the control valve without a change in load, the fl ow changes immediately. In any large system, pressure fl uctua-tions are unavoidable as pumps go on- and off-line and change speed, and as loads and fl ows vary at other locations. Figure 6 is a good example of this.
It takes time for a conventional control valve actuator to return the valve to the right position for the load. Pressure dependent control valves hunt constantly as the hydraulic profi le changes, even with sophisticated controls.
Valve and Actuator Capability Valve and Actuator Capability
As Figure 7 illustrates, when ∆T is low, the differential pressure T is low, the differential pressure T close to the plant is higher than anticipated. Many conventional control valves and their actuators lack the capability to handle
the extra differential pressure and may begin to operate like on-off valves. With the added pressure, balancing valves only function to clip maximum fl ow and do nothing to optimize coil performance at part load.
Pressure Independent Chilled Water Control
With pressure independent control, the fl ow rate at any given cooling coil changes only when the load changes. Typically, the air handler leaving air temperature relative to setpoint is used to control the actuator to position the valve.
Figure 8 illustrates pressure independent modulating two-way control valves. Mechanically, the piston and spring in the valve function to maintain a constant differential pressure across the control surface within the valve (P1 – P2). To change the fl ow rate, the stem must be rotated. A change in the pressure upstream or downstream of the valve will not change the fl ow.
Applying pressure independent modulating two-way control valves eliminates the effect of pressure fl uctuations on the fl ow rate through cooling coils. It also creates a system robust to un-certainty or changes in the hydraulic profi le. High ∆T at all load T at all load T conditions is the result.4,5
An ideal building confi guration ensures that high ∆T is achieved at each coil in the building and eliminates the need for building pumps to be run when not required (Figure 2
pumps to be run when not required (Figure 2
pumps to be run when not required ( , Scheme G). If the building is not in a hydraulically remote location, the building pump can be removed so that only a supply and return pipe is necessary to provide chilled water (Figure 2
necessary to provide chilled water (Figure 2
necessary to provide chilled water ( , Scheme F). By using pressure independent modulating control valves and eliminating return water blending with supply, low CHWST is maintained, ∆T is maximized, and the coil’s capacity T is maximized, and the coil’s capacity T is increased.
It is easy to conceptualize how raising the ∆T and minimizing T and minimizing T the use of building pumps, crossover bridges, and deny valves saves pump energy. Variable speed fan energy savings also may be realized by enabling the system to maintain low leaving air temperature off the coil.
Energy Savings Estimate Energy Savings Estimate
The following example is intended to illustrate the economic impact of the pump and fan energy savings possible in a single 200 180 160 140 120 100 80 60 40 System P ressure (ft) Distance Low 6° Low 8° Design 12°
Expansion Tank Pressure 20 ft drop across coils and control valves
Figure 7: Hydraulic gradient at design through range of ∆T
12 ft 12 ft 46.5°F 55°F LAT 62°F LAT 42°F BEFORE AFTER 60°F 54°F T
Figure 9: Change from decoupled building with return water temperature control to directly connected building with pressure independent modulating two-way control valves.
Schemes F and G). Alternatively, as shown in Table 2, if the building pump is run, the least energy consumption occurs at the highest ∆T. T. T
Note that with ∆T at 7.5°F (4.2°C) the pump motor is run-T at 7.5°F (4.2°C) the pump motor is run-T ning at maximum power despite the installed VFD. Only by minimizing both the fl ow rate and pressure drop by raising ∆T will the VFD perform as intended to save energy.
If a conventional system is designed with circuit setters, these may limit total fl ow through the building to 1,120 gpm (71 L/s). However, if CHWST is rising due to return water blending with supply, coil capacity is reduced and it will need more fl ow to satisfy the same load. The resulting effect is a rise in supply air temperature and perhaps an increase in airside energy or reduction in occupant comfort.
Locating the differential pressure sensor across the hy-draulically most remote control valve has an impact on pump energy consumption. Since an installed cooling coil is merely a fi xed pipe, the differential pressure can be placed across the hydraulically most removed valve (in lieu of the entire circuit) to save pump energy at part load.
AHU Fan Energy AHU Fan Energy
For this portion of the analysis, we will look only at the air handler fan energy that is consumed at peak load in a typical 100% outside air building as the supply air temp-erature rises. cfm = 12,000 × tonssensible 1.085 × (EAT – LAT) 1.085 × (EAT – LAT) bhpfan bhpfan
bhp = cfmactualactualactual × PD × PD × PD × PDfanfan
6356 × bhp 6356 × bhpfan 6356 × fan bhpfan bhp 6356 × bhpfan bhp ηfan ηfan η kWfan_motor kWfan_motor kW = 0.746 × bHP × bHP × bHPfanfan ηmotor
building that converts from return water temperature control with conventional control valves to direct connect with high quality pressure independent modulating control valves. The example also factors in an incentive paid by the thermal utility for raising ∆T across the building. T across the building. T Figure 9 is an illustration of the conversion.
Building Mechanical Confi guration Building Mechanical Confi guration
The building considered is a 100% outside air facility with three large variable speed air-handling units designed for 80,000 cfm (37 760 L/s) and 6 in. H2O (1.5 kPa) pressure drop each. Cooling coils are designed for 42°F (5.5°C) CHWST and a 15°F (8.4°C) ∆T. The pump serving the building is capable T. The pump serving the building is capable T of delivering 1,120 gpm (71 L/s) at 27 ft (81 kPa) of head. Design cooling load is 700 tons (2462 kW) (400 tons sensible [1407 kW]). The estimate considers the effect of rising supply air temperature and declining ∆T on a part load day. T on a part load day. T Building Pump Energy
Building Pump Energy
With a total design cooling load of 700 tons (2462 kW), de-sign fl ow for the building at 15°F (8.4°C) ∆T is 1,120 gpm (71 T is 1,120 gpm (71 T L/s). A VFD controls pump speed to maintain 12 ft (36 kPa) across the hydraulically most remote coil control valve.
∆T = T = T CHWRT – CHWST gpm = tons ×24 ∆T bhppump bhppump bhp = gpm ×PD×PD×PDpumppump 3960 × bhp 3960 × bhppump 3960 × pump bhppump bhp 3960 × bhppump bhp ηpump ηpump η kWpump_motor kWpump_motor kW = 0.746 × bhp × bhp × bhppumppump ηmotor
Actual Building Load Actual Building Load
350 tons (1231 kW) total cooling load (280 tons sensible [985 kW])
40°F (4°C): design CHWST 55°F (13°C): design CHWRT. Variable Speed Pump Details Variable Speed Pump Details
One 1,120 gpm (71 L/s) pump
7 ft (21 kPa) drop each through the supply and return (at design fl ow)
8 ft (24 kPa) drop across coil (at design fl ow) 12 ft (36 kPa) drop across control valve (always) 85% pump effi ciency (assumed)
90% pump motor effi ciency (assumed)
Using the above equations, the pump design power is: bhp = 9 (6.7 kW).
If the system pump is providing enough pressure, the best option is to serve the load without the building pump running. At locations closer to the central plant, the building pump is removed or a bypass is installed around the pump (Figure 2,
Actual Building Load Actual Building Load
350 tons (1231 kW) total cooling load (280 tons sensible [985 kW])
75°F (23.9°C): space temperature setpoint 55°F (12.8°C): leaving air temperature setpoint. Variable Speed Fan Details
Variable Speed Fan Details
Three 80,000 cfm (37 760 L/s) air handlers
6 in. H2O (1.5 kPa) total pressure drop at 80,000 cfm (37 760 L/s)
1 in. H2O (0.25 kPa) static pressure setpoint 75% fan effi ciency (assumed)
90% fan motor effi ciency (assumed)
Using the above equations, the fan design power is: bhp = 100.6 (75 kW) each.
For the analysis below, the static pressure setpoint is 1 in. H2O with 6 in. H2O (1.5 kPa) total pressure drop at 80,000 cfm (37 760 L/s) through each air handler. To simplify analysis, fan effi ciency was assumed to be 75% and fan motor effi ciency was assumed to be 90%. Table 3 illustrates airfl ow and motor power requirements in the building at part load with rising supply air temperature.
In many chilled water system applications with return water temperature control, rising CHWST prevents the system from achieving low supply air temperature (SAT). The price paid is in additional fan horsepower and thermal comfort. In the case illustrated previously, if the SAT could be reduced from 62°F to 55°F (16.7°C to 12.8°C) without violating minimum airfl ow requirements, fan energy could be reduced from 244.4 to 83.5 kW, a 65.8% reduction.
Keys to Reduced Building Fan and Pump Energy Keys to Reduced Building Fan and Pump Energy
It is the combination of low supply air temperature and high ∆T that will minimize pump and fan energy consumption in T that will minimize pump and fan energy consumption in T buildings. If return water is blended with supply in any applica-tion, coil capacity is reduced, ∆T declines, and gpm/ton rises. T declines, and gpm/ton rises. T Variable speed fan energy also will rise if the intended SAT rises above setpoint.
Figure 10 illustrates the adverse effect of return water temperature control in an actual system. In this laboratory, the CHWST to the building is 45°F (7.2°C), but it rises in the building to as high as 57°F (13.9°C). SAT rises to 62°F (16.7°C)
resulting in excess airfl ow, humidity concerns and excess fan energy consumption. ∆T within the building is far below the T within the building is far below the T 12°F (6.7°C) design ∆T for the coils.T for the coils.T
In contrast, Figure 11 illustrates performance in a system designed as shown in Figure 8. No return water blending with supply is permitted. Pressure independent modulating control valves on coils are used. CHWST remains at 40°F (4.4°C) per-mitting the SAT to be reduced to 55°F (12.8°C). Airfl ow in this case can be minimized to save fan energy. ∆T exceeds the 15°F T exceeds the 15°F T (8.4°C) ∆T for the coil at all load conditions, permitting chillers T for the coil at all load conditions, permitting chillers T in the plant to be fully loaded and pump energy reduced.
Total cooling energy costs in a customer’s facility are a combination of chilled water rates from the utility and the electricity consumed operating pumps and fans in the buildings. As an illustration, Figure 11 shows the performance improve-ment achieved when a decoupled building with return water temperature control is converted to a direct connect building with pressure independent modulating control valves at the cooling coils.
Table 4 is a summary of expected customer energy cost sav-ings for three 24/7 buildsav-ings with the design cooling character-istics described previously but located in three different cities. Electricity rates are assumed to be $0.07/kWh. Incentive chilled water rates, as described in Figure 1, are applied.
Conclusions
Return water temperature control is one example of a design strategy that increases ∆T across the building to reduce chilled T across the building to reduce chilled T water fl ow at the plant and in the main distribution. Unfor-tunately, the benefi t to the central plant and distribution can compromise building performance.
With the growing application of high quality pressure inde-pendent modulating control valves, no need exists to decouple buildings just to address low ∆T issues at coils. Despite unavoid-T issues at coils. Despite unavoid-T able system pressure fl uctuations or the evolution in the hydrau-lic profi le over time, these valves help generate the maximum possible ∆T to minimize fl ow requirements at any load.T to minimize fl ow requirements at any load.T
If the district cooling industry hopes to further improve per-formance over self-generated cooling, it is vital that we choose components that maximize chilled water ∆T performance at cooling coils. Incentive chilled water rates help foster invest-100 90 80 70 60 50 40 30 20 10 0 Temperature (°F) May 21, 2004 100 90 80 70 60 50 40 30 20 10 0 Temperature (°F) June 30, 2002
Figure 10 (left): Cooling coil performance with return water temperature control. Figure 11 (right): Cooling coil performance with pres-sure independent modulating control valves.
ments in equipment to optimize performance. High ∆T in the distribution and at the central plant improves diversity and increases available system capacity permitting the system to cool more customers with less equipment and less energy.6 This should translate into more cost-effec-tive service.
Typical benefi ts customers can expect with directly connected buildings and pressure independent modulating control valves include less fan energy consumption, less building equipment, simpler system operation, reduced capital and operating expenses, and better humidity control.
References
1. Rishel, J. 2004. “System Optimization with Hydraulic Gradi-ent Analysis and Pressure IndependGradi-ent Control Valves.” University of Wisconsin Chilled Water Plant Seminar.
2. Moe, E.M. 2005. “Win-Win Control Strategies for District Cooling Customers.” Presented at the International District Energy Association Campus Conference.
3. “Control Valve Selection.” 2005. Sponsoring Technical Com-mittee 1.4, Control Theory and Application. Project Work Statement Version 1, Feb. 6.
4. “Control valves exceed design engineers expectations.” 2003.
The Chief Engineer (12). The Chief Engineer (12). The Chief Engineer
Seattle Boston Dallas
50% Load (tons) 350 350 350 Sensible Load (tons) 280 280 280 Equivalent Time at 50% Load (hours/year) 1,747 3,119 7,124 Pump Power Savings (kW)—7.5°F to 18°F ∆T 5.8 T 5.8 T 5.8 5.8 5.85.8 Fan Power Savings (kW)—60°F to 55°F LAT 85.2 85.2 85.2 Annual Electricity Savings (kWh) 158,977 283,829 648,284 Electricity Consumption ($/kWh) $0.07 $0.07 $0.07 Electricity Savings (per year) $11,128 $19,868 $45,380 Total Chilled Water Cooling (ton-hours/year) 611,450 1,091,650 2,493,400 Incentive to Increase ∆T
At Building Interface from 10°F to >15°F $0.0461 $0.0461 $0.0461 Reduction in CHW Expense (per year) $28,188 $50,325 $114,946 Total Customer Energy $39,316 $70,193 $160,326 Cost Savings (per year)
Table 4: Annual cooling energy cost savings for customers.
5. Borer, E. and Schwartz, J. 2005. “High Marks for Chilled-Water System: Princeton Upgrades and Expands.” District Energy, First Quarter.
6. Skoglund, P.K. 2003. “Control Your Chilled Water—Save Energy / Increase Capacity.” International District Energy Association Confer-ence Presentation.
