ASHRAE OR-16-C012-2016 A Deeper Look at Modern Heating System Design.pdf

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1、Omar Hawit is a Professional Engineer and Associate Principal at Westlake Reed Leskosky, Washington, DC. Christopher Wilson is a Professional Engineer and Senior Mechanical Engineer at Westlake Reed Leskosky, Washington, DC. Trevor Jaffe is a Professional Engineer and Mechanical Engineer at Westlake

2、 Reed Leskosky, Washington, DC. A Deeper Look at Modern Heating System Design Omar Hawit, PE Trevor Jaffe, PE Christopher Wilson, PE Member ASHRAE Member ASHRAE Member ASHRAE ABSTRACT The heating, ventilation, air conditioning, and refrigeration industry shows a trend toward increased adoption of co

3、ndensing boilers in heating systems, but the full impact on design and construction is complex. In design the selection of the boilers, heating coils, and operating temperatures play equally important roles in the composition of an efficient system. The failure to optimize design for a system as a w

4、hole can lead to wasted energy and money. This paper investigates the effect on overall system performance and construction cost of various heating system metrics with regards to condensing and non-condensing boilers. INTRODUCTION The paper studies the system wide effects of heating system efficienc

5、y with respect to boiler type and operating temperatures while weighing the first cost of system components against the energy cost of the system operation. The maximum efficiency of condensing boilers can significantly exceed that of non-condensing boilers at lower entering water temperatures. Howe

6、ver, in order to reach lower return water temperatures, coils must be properly selected. A standard coil selection exhibits a tendency to produce a smaller change in temperature with lower water temperatures. As a result, pumping energy increases due to the increased flow required to supply the same

7、 heat energy. The system efficiency and cost implications are reviewed at the component level. The impact of boiler types, air handling unit pre-heat coils, and variable volume air terminal air units is analyzed. The analysis culminates with a whole building energy simulation that provides a paramet

8、ric analysis of the multiple variables presented in system selection with respect to system energy use and costs. The modeled building was simulated in Washington, D.C. (ASHRAE Climate Zone 4A) in the parametric analysis. Then Duluth, MN (ASHRAE Climate Zone 7) and Los Angeles (ASHRAE Climate Zone 3

9、B) were compared. The paper concludes with a proposed range of operating temperatures. CONDENSING BOILER EFFICIENCY A typical 2,000 MBH (586 kW) condensing boiler was analyzed for thermal efficiency at a range of entering water temperatures and part load conditions. Entering water temperatures rangi

10、ng from 80F (27C) to 160F (71C) were used for analysis along with part load conditions ranging from 20% to 100%. A 20F (11C) temperature differential between entering water temperature and leaving water temperature was used. Temperature differential has little impact on the thermal efficiency of the

11、 boiler. The results indicate that thermal efficiency varies greatly at different entering water temperatures and part load OR-16-C012conditions. As entering water temperature decreases, thermal efficiency rises and in some cases can increase by 10% at a fixed part load. Dramatic increases in effici

12、ency do not start until entering water temperatures drop below 140F (60C), since this is when condensation starts to occur. Thermal efficiency also continues to rise as part load increases. Figure 1. Thermal efficiency shown with respect to part load operation. Non-condensing heating hot water boile

13、rs typically have efficiencies in the 80% to 85% range. At the high end of this range of efficiency the limit is driven by maintaining flue temperatures well above 130F (54C) to avoid condensation in the boiler or flue that would damage the heat exchanger. The highest achievable efficiency that can

14、occur without condensation forming inside the boiler is approximately 87%. Condensation will almost certainly occur immediately outside the boiler in the vent. This range is constrained at the low end by the ASHRAE 90.1 Energy Standard and DOE Energy Conservation Standards for Commercial Boilers. St

15、eam boilers will operate around 77%-79% thermal efficiency for similar reasons. The results show that condensing boilers can achieve far greater efficiencies over a larger operating range due to the ability to operate during condensing conditions. In addition, condensing boilers operate at even grea

16、ter efficiency at part load conditions, which is the predominant demand for a building. The control logic for operating multiple condensing boilers in the same heating plant should take advantage of increased thermal efficiency at part load conditions. While thermal efficiency and in turn energy sav

17、ings are far greater for condensing boilers, first cost should be taken into consideration. The typical 2,000 MBH (586 kW) high-efficiency condensing boiler is listed at a price range of $30,000-$34,000 while a typical non-condensing boiler of the same size and manufacturer is listed at a price of a

18、pproximately $22,000. At a glance, this appears to be an increase in first cost. However, there are several factors that equalize the actual installed cost. It is important to note that the size for these boilers typically refers to the gas input. Therefore, the condensing type boiler offers about 2

19、0% more capacity for the same listed size. Another way of stating this is that the project must purchase 20% more non-condensing boilers. In our example system discussed below, this is roughly the cost of a sixth additional boiler and the cost of the all associated appurtenances including piping spe

20、cialties for gas and water, OR-16-C012electrical power, and low voltage controls. Another major equalizer between the first cost of condensing and non-condensing boilers is the ability to increase the temperature differential, dT, in the heating hot water. A greater temperature delta reduces the flo

21、w rate, which in turn reduces the heating system pump size and system wide pipe, pipe specialty sizes, and pipe connections. These balance of system material costs compose roughly one third of the heating plant first costs, so a 10% cost reduction in sizing would yield about a $1,000 reduction in sy

22、stem cost for each boiler. When factoring in all first costs across the system for each boiler type, the cost of the condensing type boiler becomes much more comparable to that of the non-condensing boiler. AIR HANDLING UNIT PREHEAT COILS A typical medium office building load model was used to size

23、a variable volume air-handling unit serving a simulated 1st floor. The air-handling unit was sized to 10,100 CFM (17,160 CMH) supply air cooling at 55F (13C), 4,700 CFM (7,985 CMH) supply air heating at 55F (13C) heating and 2,500 CFM (4,250 CMH) outdoor air year round. The air-handling units cross-

24、sectional area and coil dimensions were sized to maintain a maximum of 500 FPM (2.5 MPS) through the coils at the cooling supply airflow rate. The heating hot water pre-heat coil was sized to heat 4,700 CFM (7,985 CMH) of mixed air (17.3F (-8.2C) outdoor air and 69.4F (20.8C) return air) to a leavin

25、g coil temperature of 55F (13C). Multiple heating hot water entering coil and leaving coil conditions were used to obtain flow rate, air pressure drop and water pressure drop values through the pre-heat coil. Entering water temperature values of 180F (82C), 160F (71C), 140F (60C) and 120F (49C) were

26、 used with corresponding change in temperature across the coil values of 20F (11C), 30F (17C) and 40F (22C). No variation in coil rows, fin spacing or air pressure drop was observed with varying heating hot water temperatures and the change in temperatures. As the change in water temperature across

27、the coil increased, generally the number of coil feeds decreased and corresponding number of passes increased. Water pressure drop was dependent on circuiting and did not appear to correlate to heating hot water temperatures or the change in temperatures. Table 1. Air-Handling Unit Hot Water Pre-Hea

28、t Coil Analysis # Entering Water Temperature Leaving Water Temperature Change in Water Temperature # of Rows # of Feeds # of Passes Fin Spacing Air Pressure Drop Water Pressure Drop F (C) F (C) F (C) Fins / in (Fins / cm) in W.C. (Pa) Feet (kPa) 1 180 (82) 160 (71) 20 (11) 1 16 (Half) 2 6 (2.4) 0.02

29、 (5.0) 1 (3.0) 2 180 (82) 150 (66) 30 (17) 1 16 (Half) 2 6 (2.4) 0.02 (5.0) 0.9 (2.7) 3 180 (82) 140 (60) 40 (22) 1 8 (Quarter) 4 6 (2.4) 0.02 (5.0) 2 (6.0) 4 160 (71) 140 (60) 20 (11) 1 16 (Half) 2 6 (2.4) 0.02 (5.0) 1.7 (5.1) 5 160 (71) 130 (54) 30 (17) 1 16 (Half) 2 6 (2.4) 0.02 (5.0) 1.9 (5.7) 6

30、 160 (71) 120 (49) 40 (22) 1 8 (Quarter) 4 6 (2.4) 0.02 (5.0) 1.2 (3.6) 7 140 (60) 120 (49) 20 (11) 1 16 (Half) 2 6 (2.4) 0.02 (5.0) 1 (3.0) 8 140 (60) 110 (43) 30 (17) 1 8 (Quarter) 4 6 (2.4) 0.02 (5.0) 1.6 (4.8) 9 140 (60) 100 (38) 40 (22) 1 8 (Quarter) 4 6 (2.4) 0.02 (5.0) 0.6 (1.8) 10 120 (49) 1

31、00 (38) 20 (11) 1 16 (Half) 2 6 (2.4) 0.02 (5.0) 1.4 (4.2) 11 120 (49) 90 (32) 30 (17) 1 8 (Quarter) 4 6 (2.4) 0.02 (5.0) 0.7 (2.1) 12 120 (49) 80 (27) 40 (22) 1 2 16 6 (2.4) 0.02 (5.0) 4 (12.0) For a typical variable volume air-handling unit the unit cross-sectional area is determined by the peak c

32、ooling airflow and associated coil face velocities. The pre-heat coil dimensions are determined by airflows 30-50% larger than OR-16-C012the heating airflow. This results in the pre-heat coils surface area being much larger than required for heating and minimal coil rows being required. Since the pr

33、e-heat coils are oversized, heating hot water temperatures and change in temperature values have little impact on the air pressure drop and water pressure drop values that can contribute to fan and pump energy. It should be noted that this phenomenon is specific to variable volume air-handling units

34、. For constant volume and dedicated outdoor air supply units where the heating airflow is the same as the cooling airflow, coil face velocities will be five hundred (500) FPM (2.5 MPS) and heating hot water temperatures and change in temperatures will have a large impact on fan and pump energy. Sinc

35、e heating hot water temperatures and change in temperature values have little impact on the number of rows required, it can be assumed that there is no impact on the first cost of the air-handling unit. VAV TERMINAL REHEAT COILS A typical medium office building load model was used to size a typical

36、variable air volume terminal unit for the building. The simulated zone used for sizing was a perimeter zone. The variable air volume terminal unit was sized to 1,865 CFM (3170 CMH) supply air cooling at 55F (13C) and 560 CFM (950 CMH) supply air heating at 90F (32C). Multiple heating hot water enter

37、ing coil and leaving coil conditions were used to obtain flow rate, air pressure drop and water pressure drop values through the terminal units reheat coil. As entering water temperature decreased, generally the size of the variable volume air terminal unit, number of coil rows and the air pressure

38、drop across the coil increased. There was no direct correlation observed between entering water temperature and water pressure drop across the coil or water flow rate, as these values were more dependent on coil circuiting. Table 2. Air Terminal Unit Reheat Coil Analysis # Unit Size Entering Water T

39、emperature Leaving Water Temperature Change in Water Temperature # of Rows Air Pressure Drop Water Pressure Drop Water Flow Rate F (C) F (C) F (C) IN.W.C. (Pa) Feet (kPa) GPM (CMH) 1 12 180 (82) 162.1 (72.3) 17.9 (9.9) 1 0.4 (100) 1 (3.0) 2.4 (0.54) 2 14 180 (82) 144.4 (62.4) 35.6 (19.8) 1 0.2 (50)

40、0.4 (1.2) 1.2 (0.27) 3 14 180 (82) 125.2 (51.8) 54.8 (30.4) 1 0.2 (50) 1 (3.0) 0.8 (0.18) 4 14 160 (71) 144.9 (62.7) 15.1 (8.4) 1 0.2 (50) 1.6 (4.8) 2.9 (0.66) 5 14 160 (71) 128.8 (53.8) 31.2 (17.3) 1 0.2 (50) 2.3 (6.9) 1.4 (0.32) 6 14 140 (60) 99.2 (37.3) 40.8 (22.7) 2 0.4 (100) 0.2 (0.6) 1.1 (0.25

41、) 7 14 140 (60) 89.9 (32.2) 50.1 (27.8) 2 0.5 (125) 0.4 (1.2) 0.9 (0.20) 8 14 120 (49) 103.3 (39.6) 16.7 (9.3) 2 0.4 (100) 1.1 (3.3) 2.6 (0.59) 9 14 120 (49) 95.6 (35.3) 24.4 (13.6) 2 0.5 (125) 1.4 (4.2) 1.8 (0.41) 10 14 120 (49) 85.3 (29.6) 34.7 (19.3) 3 0.7 (174) 0.3 (0.9) 1.2 (0.27) Based on the

42、results above, at the variable air volume terminal unit an increase in air pressure drop, and in result fan energy, can be associated with the HVAC system when the heating hot water entering water temperature is decreased. Air pressure drop approximately doubles in value when going from an entering

43、water temperature of 180F (82C) to 140F (60C), see Figure 2 below. OR-16-C012Figure 2. Variable volume air terminal pressure drop shown with respect to entering water temperature. It should be noted that the increase in air pressure drop across the variable air volume terminal unit in comparison to

44、the overall air pressure drop value for the system is relatively low. However, it should be factored into the overall energy equation when selecting system heating hot water temperatures. The increased fan energy use should be weighed against the gains of the heating plant efficiency. Water pressure

45、 drop across the coil, and, in result pump energy, appeared to have no correlation to entering hot water temperature and can be effectively ignored when calculating energy impacts of varying system heating hot water temperatures. Recent cost estimate for projects show that the reheat coil cost is mo

46、stly labor. An increase of terminal unit by one size yields about a 10% material cost increase. One estimate for a building with a $2.7M HVAC budget showed the total cost for terminal reheat coils appears to be about $24,000, where almost $17,000 appears to be attributes to connections piping labor

47、and piping material cost. Therefore, differences between 1-row and 2-row coils, or the difference in increasing the coil by one size, will be small relative to the total mechanical cost. The reheat coil cost as a percentage of the total heating system in this example project is less than one percent

48、. WHOLE BUILDING ENERGY SIMULATION Tranes TRACE700 whole building energy simulation tool, version 6.3.1, was used to construct a model of a typical large office building and simulate its operating energy and energy costs in Washington, D.C. Cooling and heating calculation methodologies were chosen a

49、s “RTS (ASHRAE Tables)” and “UATD”, respectively. Infiltration and outside film entries were set to “Vary with wind speed”, and terrain was set to “Center of a large city”. Simulation inputs for the building were sourced from the Pacific Northwest National Laboratorys ASHRAE 90.1 Prototype Building Modeling Specifications (henceforth referred to as the Prototype Specifications), most recently updated in March of 2014. The large office building type is 498,600 ft2 (46,322 m3) of floor area distributed OR-16-C012across tw