Table 5-6
Table 5-6 Impact on First Costs and Energy Costs of Chilled-Water Temperature Difference (Assuming Constant Chilled-Water Supply Temperature)
T
Low High
Typical range 8°F 25°F
First-cost impact Smaller coil
Smaller pipe Smaller pump Smaller pump motor
Energy-cost impact Lower fan energy Lower pump energy
shows the typical range of CHW temperature difference (com- monly referred to as delta-T or T) and the general impact on energy usage and first costs. The table shows that there are significant benefits to increas- ing T from a first-cost standpoint, and there may be energy cost savings as well, depending on the relative size of the fan energy increase (due to increased air-side pressure drop with deeper CHW coils) versus pump energy decrease as T increases. Chiller energy usage is largely unaffected by T
for a given CHW supply temperature. The leaving CHW temperature drives the evaporator temperature, which in turn drives chiller efficiency; entering water temperature has almost no impact on efficiency.
Intuitively, one might think that fan energy would dominate in the energy balance between fan and CHW pump because fan energy is so much larger than pump energy annually, and the fan sees the coil pressure drop under all conditions, while the CHW pump typically only runs in warmer weather (assuming the system has an air-side economizer). However, detailed analysis has shown that not to be the case: the impact on the air side of the system is sel- dom significant. Table 5-7
Table 5-7 Typical Coil Performance Versus Chilled-Water Temperature Difference
Chilled-WaterT, °F 5.5 7.2 8.9 10.6 12.2 14
Coil water pressure drop, ft H2O 7.2 4.2 2.8 2.5 2.0 1.4
Coil air-side pressure drop, in. H2O 12.2 12.7 13.2 15.2 16.0 19.8
Rows 6 6 6 8 8 8
Fins per in. (fpi) 2.9 3.3 3.7 3.0 3.4 4.6
Cooling coil pressure air- and water-side drops were determined from a manufacturer’s AHRI-certified selection program assum- ing 500 fpm coil face velocity, smooth tubes, maximum 12 fpi fin spacing, 43°F CHW supply temperature, 78°F/63°F entering air temperature, and 53°F leaving air temperature.
shows a typical cooling coil’s performance over a range of CHWTs. While the example in the table will not be true of all appli- cations, it does suggest that air-side pressure will not increase very much as CHWT rises, while water-side pressure drop falls significantly. For VAV sys- tems, the impact on annual fan energy is even less significant because any full- load air-side pressure drop penalty will fall rapidly as airflow decreases.
Figure 5-9 shows the impact of CHWT on energy usage for a typical Oak- land, CA, office building served by a VAV air-distribution system with VFD and an air-side economizer. Fan energy rises only slightly as T increases. If pipe size is left unchanged asT increases, CHW pump energy will fall substantially due to reduced flow and reduced piping losses. In real applications, pipe sizes are generally reduced to decrease first costs, but pump energy will still fall due to reduced flow rates and reduced coil and evaporator pressure drops, although not as dramatically as in Figure 5-9.
Reducing CHW temperature can eliminate the fan energy penalty. Figure 5-10 shows the same system as Figure 5-9, but instead of holding CHW temperature constant and increasing coil size to increaseT, CHW temperature is lowered and coil size is held constant to keep air-side pressure drop (and
therefore fan energy) constant as T increases. Dropping CHW temperature
increases chiller energy, but pump energy savings more than make up the dif- ference. As with Figure 5-9, the pump energy shown in Figure 5-10 assumes that pipe sizes remain constant, which is not always the case.
Figure 5-9 Typical annual energy usage versus CHW T with a constant CHW supply tem- perature and constant pipe sizes.
Source: Taylor 2011b.
Figure 5-10 Typical annual energy usage with coils selected for constant air-side pressure drop.
Source: Taylor 2011b.
Table 5-8 compares three cooling coils with 4, 6, and 8 rows that result in about 10°F, 18°F, and 25°F T, respectively. Pipe sizes were selected from Table 5-4 assuming an acoustically sensitive location with about 2000 h/yr of operation. Coil first costs were obtained from the manufacturer’s representative
Table 5-8 Cooling Coil and Associated Piping Costs
(For 20,000 cfm coil sized at 500 fpm, 42°F CHW supply temperature, 78°F entering dry-bulb temperature, 62°F entering wet-bulb temperature, and 53°F leaving dry-bulb temperature)
Coil Piping Fins per in. Rows Air Pressure Drop, in. H2O FluidT, °F Fluid Flow,gpm Fluid Pressure Drop, ft H2O
Coil Cost Pipe Size, in. Coil Connection Total Cost 10 4 17.8 10.1 118.7 9.1 $3598 3 $4551 $8149 11 6 16.51 18.2 66.0 7.6 $4845 2.5 $3581 $8426 10 8 20.32 24.9 47.0 5.7 $5956 2 $2101 $8057
and piping costs (including typical valve train and 20 ft of branch piping) were obtained from the LCCA spreadsheet piping cost data. Table 5-8 shows that the added cost of the deeper coil is more than offset by the savings in the cost of piping the coil. And there are additional first-cost savings from the reduced piping main, pump, pump motor, and VFD sizes.
So, increasing T reduces both first costs and energy costs. Clearly life- cycle costs will be lower the higher theT. We were unable in our analysis to find a point where the negative impact on fan energy or coil costs caused life- cycle costs to start to rise with increasingT; within the range of our analysis (up to 25°FT), higher T was always better. Energy savings from high T are even greater with systems that have WSEs or CHW TES. To reiterate: our anal- ysis suggests that it never makes sense to use the traditional 10°F or 12°FTs that are commonly used in standard practice.
As T is increased, eventually the ever-deepening coil will run into the ASHRAE Standard 62.1 coil pressure drop limit (2016a). ASHRAE Standard 62.1 uses dry coil pressure as a surrogate for the cleanability of the coil. Section 5.11.12 of the standard requires that dry coil pressure drop at 500 fpm face velocity must not exceed 0.75 in. This is roughly the pressure drop of an 8 row, 12 fpi coil (depending on the details of the fin and tube design).
So, the design procedure for selecting CHW coils is simple: rather than arbitrarily selecting CHW temperatures and then selecting coils that deliver those temperatures, reverse the logic: always use the “biggest” (highest effec- tiveness) coil available without exceeding the ASHRAE Standard 62.1 pres- sure drop limits and let the CHWT be determined by the coil and design air conditions.
Based on this logic, the recommended procedure for sizing CHW coils and selecting CHW design temperatures is as follows. The intent of this procedure is to achieve all of the piping first cost savings resulting from a high T but with as warm a CHW temperature as possible to improve chiller efficiency.
1. Calculate the CHW flow rate for all coils assuming a 25°FT.1
2. Pick primary pipe sizes (at pumps, headers, main risers, and main branch lines) in the critical circuit (that which determines pump head) using Table 5-4 or the LCCA spreadsheet.
3. With pipe sizes selected, use Table 5-4 or work backwards using the LCCA spreadsheet to find the maximum flow for each pipe size, and then recalcu- late the T in each pipe using these flow rates. This is the minimum aver- ageT for this leg of the circuit.
4. Use the coil manufacturer’s selection program to find the maximum coil size that complies with the ASHRAE Standard 62.1 cleanability limit, typ- ically 8 row/12 fpi.2 Use this for all coils so that T is maximized. (For some smaller fan-coils, 8 row coils may not be an option. If so, use the larg- est coil available but no less than 6 rows. If that is not an option with the selected manufacturer, find another manufacturer.)
5. With the coil manufacturer’s selection program, iterate on coil selections to determine the CHW supply temperature that results in the selected T on
average for each leg of the critical circuit, starting with the coil at the end of
the circuit and working back to the plant. It is not necessary that allTs be the same (and in fact they definitely will not be the same with this approach), just that the flow through the circuit equals the maximum flow determined in
Step 3. The recommended minimum CHW supply temperature is 42°F.2If
this minimum is reached in any leg and the flow exceeds the maximum, then the process must be started over with a smallerT assumption in Step 1. 6. The lowest required CHW supply temperature for any coil in the circuit is
then the design temperature.
7. Determine actual flow andT in other coils in other circuits using the coil selection program with this design CHW supply temperature, again maxi- mizing coil size within ASHRAE Standard 62.1 limits (e.g., 8 rows, 12 fpi) and letting the program determine return water temperature.
8. The plant flow is that calculated using the diversified (concurrent) load and the gpm-weighted average return water temperature of all coils.
1. Some engineers may be concerned that a 25°FT is nonconservative and reduces future flexibility for load changes. Designing around largeTs results in large coils and small pipes and pumps. Designing around smallTs results in the opposite, small coils and large pipes and pumps. Both are equally forgiving with respect to possible coil load changes—one is no more conservative than the other. If excess capacity is desired for future flexibility, it should be explicitly built into the design rather than relying on acci- dental flexibility from design parameters.
2. In our analyses, the lowest CHW supply temperature resulting from this technique was about 42°F; we do not know if lower CHW temperatures will start to affect the life- cycle cost due to reduced chiller efficiency. So, unless the designer performs additional LCCA, we suggest limiting the design CHW supply temperature to no colder than 42°F. For most applications, this low temperature will not be required to achieve the target 25°FT. Limiting the supply temperature to 42°F also provides some conservatism in the design; should there be a miscalculation in loads or unexpectedly high loads at a cer- tain coil, CHW temperature can be lowered below 42°F to increase coil capacity, although with a resultant loss in overall chiller capacity and efficiency.
If this procedure seems too long and complicated, here is a shortcut proce- dure: skip Steps 1 to 5 and just assume a CHW supply temperature of 42°F in Step 6. This will provide basically the same result except that the design CHW temperature may be lower than it needed to be to achieve the pipe size savings from high T, so the chiller design efficiency may be worse than it needed to be. The energy impact will be minimal, however, because CHW temperature should be aggressively reset, as discussed in Chapter 7. This simplified approach also results in a somewhat lower CHW flow rate, so pump size and power will be reduced.