SECTION 16 HEATING, VENTILATING, AND AIR CONDITIONING

Transcription

1 Source: HANDBOOK OF MECHANICAL ENGINEERING CALCULATIONS SECTION 16 HEATING, VENTILATING, AND AIR CONDITIONING ECONOMICS OF INTERIOR CLIMATE CONTROL 16.2 Equations for Heating, Ventilation, and Air-Conditioning Calculations 16.2 Determining Cooling-Tower Fan Horsepower Requirements Choosing an Ice Storage System for Facility Cooling Annual Heating and Cooling Energy Loads and Costs Heat Recovery Using a Run-Around System of Energy Transfer Rotary Heat Exchanger Energy Savings Savings from Hot-Deck Temperature Reset Air-to-Air Heat Exchanger Performance Steam and Hot-Water Heating Capacity Requirements for Buildings Heating Steam Required for Specialized Rooms Determining Carbon Dioxide Buildup in Occupied Spaces Computing Bypass-Air Quantity and Dehumidifier Exit Conditions Determination of Excessive Vibration Potential in Motor-Driven Fan Power Input Required by Centrifugal Compressor Evaporation of Moisture from Open Tanks Checking Fan and Pump Performance from Motor Data Choice of Air-Bubble Enclosure for Known Usage Sizing Hydronic-System Expansion Tanks SYSTEM ANALYSIS AND EQUIPMENT SELECTION Building or Structure Heat-Loss Determination Heating-System Selection and Analysis Required Capacity of a Unit Heater Steam Consumption of Heating Apparatus Selection of Air Heating Coils Radiant-Heating-Panel Choice and Sizing Snow-Melting Heating-Panel Choice and Sizing Heat Recovery from Lighting Systems for Space Heating Air-Conditioning-System Heat-Load Determination General Method Air-Conditioning-System Heat-Load Determination Numerical Computation Air-Conditioning System Cooling-Coil Selection Mixing of Two Airstreams Selection of an Air-Conditioning System for a Known Load Sizing Low-Velocity Air-Conditioning- Systems Ducts Equal Friction Method Sizing Low-Velocity Air-Conditioning Ducts Static-Regain Method Humidifier Selection for Desire Atmospheric Conditions Use of the Psychrometric Chart in Air- Conditioning Calculations Designing High-Velocity Air- Conditioning Ducts Air-Conditioning-System Outlet- and Return-Grille Selection Selecting Roof Ventilators for Buildings Vibration-Isolator Selection for an Air Conditioner Selection of Noise-Reduction Materials Choosing Door and Window Air Curtains for Various Applications

2 16.2 ENVIRONMENTAL CONTROL Economics of Interior Climate Control EQUATIONS FOR HEATING, VENTILATION, AND AIR-CONDITIONING CALCULATIONS A variety of calculation procedures are used in designing heating, ventilating, and air-conditioning systems. To help save time for design and application engineers, technicians, and consulting engineers, some 75 design equations are presented at the start of this section of the handbook. These equations are used in both manual and computer-aided design (CAD) applications. And since this handbook is designed for worldwide use, the first group of equations presents USCS and SI versions to allow easy comparisons of the results. Abbreviations used in the equations follow this presentation. Btu min HS 1.08 CFM T (1) 3 h ft F kj min HSM CMM T M (2) 3 h m C Btu min lb DA HL 0.68 CFM W (3) 3 h ft grho 2 kj min kg DA HLM 177,734.8 CMM W M (4) 3 h m kgho 2 lb min HT 4.5 CFM h (5) 3 h ft kg min HTM CMM h M (6) h m 3 H H H (7) T S L H H H (8) TM SM LM Btu min H 500 GPM T (9) h gal F kj min HM LPM T M (10) h L C AC CFM 60 min/h HR VOLUME AC HR M CMM 60 min/h VOLUME M (11) (12)

3 HEATING, VENTILATING, AND AIR CONDITIONING 16.3 F 32 C (13) 1.8 F 1.8 C 32 (14) where H S sensible heat, Btu/h H SM sensible heat, kj/h H L latent heat, Btu/h H LM latent heat, kj/h H T total heat, Btu/h H TM total heat, kj/h H total heat, Btu/h H M total heat, kj/h T temperature difference, F T M temperature difference, C W humidity ratio difference, gr H 2 O/lb DA W M humidity ratio difference, kg H 2 O/kg DA h enthalpy difference, Btu/lb DA h enthalpy difference, kj/lb DA CFM airflow rate, ft 3 /min CMM airflow rate, m 3 /min GPM water flow rate, gal/min LPM water flow rate, L/min AC/HR air change rate per hour, English AC/HR M air change rate per hour, SI AC/HR AC/HR M VOLUME space volume, ft 3 VOLUME M space volume, m 3 kj/h Btu/h CMM CFM LPM GPM kj/lb Btu/lb m ft m 2 ft m 3 ft kg lb GPM 500 lb steam/h 1.0 lb steam/h GPM 1.0 lb H 2 /h 1.0 lb steam/h kg/m 3 lb/ft (density) m 3 /kg ft 3 /lb specific volume kg H 2 O/kg DA gr H 2 O/lb DA/7000 lb H 2 O/lb DA Steam Pipe Pressure Drop and Flow Rate Equations W (1 3.6/ID) P (15) 3600 D ID 5 P D ID W 5 60 (16) (1 3.6/ID)

4 16.4 ENVIRONMENTAL CONTROL W VAINCHESD 60VAFEETD (17) 2.4W W V (18) A D 60A D INCHES FEET where P pressure drop per 100 ft of pipe (psig/100 ft) W steam flow rate, lb/h ID actual inside diameter of pipe, in D average density of steam at system pressure, lb/ft 3 V velocity of steam in pipe, ft/min A INCHES actual cross-sectional area of pipe, in 2 A FEET actual cross-sectional area of pipe, ft 2 Condensate Piping Equations HSSS HSCR FS 100 (19) H LCR FS WCR W (20) 100 where FS flash steam, % H SSS sensible heat at steam supply pressure, Btu/lb H SCR sensible heat at condensate return pressure, Btu/lb H LCR latent heat at condensate return pressure, Btu/lb W steam flow rate, lb/h W CR condensate flow based on percentage of flash steam created during condensing process, lb/ h. Use this flow rate in steam equations above to determine condensate return pipe size. HVAC Efficiency Equations BTU OUTPUT EER COP (21) BTU INPUT BTU OUTPUT EER (22) WATTS INPUT Turndown ratio maximum firing rate minimum firing rate (that is 5 1, 10 1, 25 1) GROSS BTU OUTPUT OVERALL THERMAL EFF GROSS BTU INPUT 100% (23) BTU INPUT BTU STACK LOSS COMBUSTION EFF BTU INPUT 100% (24)

5 HEATING, VENTILATING, AND AIR CONDITIONING 16.5 Overall thermal efficiency range 75% 90% Combustion efficiency range 85% 95% Equations for HVAC Equipment Room Ventilation For completely enclosed equipment rooms: where CFM exhaust airflow rate required, ft 3 /min G mass of refrigerant of largest system, lb For partially enclosed equipment rooms: where FA ventilation-free opening area, ft 2 G mass of refrigerant of largest system, lb 0.5 CFM 100 G (25) 0.5 FA G (26) Psychrometric Equations The following equations are from Carrier Corporation publications.* These equations cover air mixing, cooling loads, sensible heat factor, bypass factor, temperature at the apparatus, supply air temperature, air quantity, and determination of air constants. Abbreviations and symbols for the equations are given below. Abbreviations adp BF (BF) (OALH) (BF) (OASH) (BF) (OATH) Btu/ h cfm, ft 3 /min db dp ERLH ERSH ERTH ESHF F fpm, ft/ min gpm, gal/ min apparatus dew point bypass factor bypassed outdoor air latent heat bypassed outdoor air sensible heat bypassed outdoor air total heat British thermal units per hour cubic feet per minute dry-bulb dew point effective room latent heat effective room sensible heat effective room total heat effective sensible heat factor degrees Fahrenheit feet per minute gallons per minute *Handbook of Air-Conditioning System Design, McGraw-Hill, New York, various dates.

6 16.6 ENVIRONMENTAL CONTROL Symbols gr/ lb GSHF GTH GTHS OALH OASH OATH rh RLH RLHS RSH RSHF RSHS RTH Sat Eff SHF TLH TSH wb cfm ba cfm da cfm oa cfm ra cfm sa h h adp h cs h ea h la h m h oa hr m h sa t t adp t edb t es grains per pound grand sensible heat factor grand total heat grand total heat supplement outdoor air latent heat outdoor air sensible heat outdoor air total heat relative humidity room latent heat room latent heat supplement room sensible heat room sensible heat factor room sensible heat supplement room total heat saturation efficiency of sprays sensible heat factor total latent heat total sensible heat wet-bulb bypassed air quantity around apparatus dehumidified air quantity outdoor air quantity return air quantity supply air quantity specific enthalpy apparatus dew point enthalpy effective surface temperature enthalpy entering air enthalpy leaving air enthalpy mixture of outdoor and return air enthalpy outdoor air enthalpy room air enthalpy supply air enthalpy temperature apparatus dew point temperature entering dry-bulb temperature effective surface temperature

7 HEATING, VENTILATING, AND AIR CONDITIONING 16.7 t ew t ewb t ldb t lw t lwb t m t oa t rm t sa W W adp W ea W es W la W m W oa W rm W sa entering water temperature entering wet-bulb temperature leaving dry-bulb temperature leaving water temperature leaving wet-bulb temperature mixture of outdoor and return air drybulb temperature outdoor air dry-bulb temperature room dry-bulb temperature supply air dry-bulb temperature moisture content or specific humidity apparatus dew point moisture content entering air moisture content effective surface temperature moisture content leaving air moisture content mixture of outdoor and return air moisture content outdoor air moisture content room moisture content supply air moisture content Air Mixing Equations (Outdoor and Return Air) cfmoa toa cfmra trm tm (27) cfm sa (cfmoa h oa) (cfmra h rm) hm (28) cfm sa (cfmoa W oa) (cfmra W rm) Wm (29) cfm sa Cooling Load Equations ERSH RSH (BF)(OASH) RSHS* (30) ERLH RLH (BF)(OALH) RLHS* (31) *RSHS, RLHS, and GTHS are supplementary loads due to duct heat gain, duct leakage loss, fan and pump horsepower gains, etc.

8 16.8 ENVIRONMENTAL CONTROL ERTH ERLH ERSH (32) TSH RSH OASH RSHS* (33) TLH RLH OALH RLHS* (34) GTH TSH TLH GTHS* (35) RSH 1.08 cfm (t t ) (36) sa rm sa RLH 0.68 cfm (W W ) (37) sa rm sa RTH 4.45 cfm (h h ) (38) sa rm sa RTH RSH RLH (39) OASH 1.08 cfm (t t ) (40) oa oa rm OALH 0.68 cfm (W W ) (41) oa oa rm OATH 4.45 cfm (h h ) (42) oa oa rm OATH OASH OALH (43) (BF)(OATH) (BF)(OASH) (BF)(OALH) (44) ERSH 1.08 cfm (t t )(1 BF) (45) da rm adp ERLH 0.68 cfm (W W )(1 BF) (46) da rm adp ERTH 4.45 cfm (h h )(1 BF) (47) da rm adp TSH 1.08 cfm (t t )* (48) da edb ldb TLH 0.68 cfm (W W )* (49) da ea la GTH 4.45 cfm (h h )* (50) da ea la Sensible Heat Factor Equations RSH RSH RSHF RSH RLH RTH (51) ERSH ERSH ESHF ERSH ERLH ERTH (52) TSH TSH GSHF TSH TLH GTH (53) *RSHS, RLHS, and GTHS are supplementary loads due to duct heat gain, duct leakage loss, fan and pump horsepower gains, etc. See below for the derivation of these air constants. When no air is to be physically bypassed around the conditioning apparatus, cfm da cfm sa.

9 HEATING, VENTILATING, AND AIR CONDITIONING 16.9 Bypass Factor Equations tldb tadp tedb tldb BF 1 BF (54) t t t t edb adp edb adp Wla Wadp Wea Wla BF 1 BF (55) W W W W ea adp ea adp hla hadp hea hla BF 1 BF (56) h h h h ea adp ea adp Temperature Equations at the Apparatus (cfmoa t oa) (cfmra t rm) t edb* (57) cfm sa t t BF(t t ) (58) ldb adp edb adp Both t ewb and t lwb correspond to the calculated values of h ea and h la on the psychrometric chart. (cfmoa h oa) (cfmra h rm) h ea* (59) cfm sa h h BF(h h ) (60) la adp ea adp Temperature Equations for Supply Air RSH tsa trm (61) 1.08cfm sa Air Quantity Equations ERSH cfmda (62) 1.08(1 BF)(t t ) rm ERLH cfmda (63) 0.68(1 BF)(W W ) rm ERTH cfmda (64) 4.45(1 BF)(h h ) rm adp adp adp *When t m, W m, and h m are equal to the entering conditions at the cooling apparatus, they may be substituted for t edb, W ea, and h ea, respectively. See footnote on page 16.9.

10 16.10 ENVIRONMENTAL CONTROL TSH cfm da* (65) 1.08(t t ) edb ldb TLH cfm da* (66) 0.68(W W ) ea la GTH cfm da* (67) 4.45(h h ) ea la RSH cfmsa (68) 1.08(t t ) rm sa RLH cfmsa (69) 0.68(W W ) rm sa RTH cfmsa (70) 4.45(h h ) rm sa cfm cfm cfm (71) ba sa da Note: cfm da will be less than cfm sa only when air is physically bypassed around the conditioning apparatus. cfm cfm cfm (72) sa oa ra Derivation of Air Constants (73) 13.5 where specific heat of moist air at 70 F db and 50% rh, Btu/( F lb DA) 60 min/h 13.5 specific volume of moist air at 70 F db and 50% rh where 60 min/h 13.5 specific volume of moist air at 70 F db and 50% rh 1076 average heat removal required to condense 1 db water vapor from the room air 7000 gr/lb *See footnote on page 16.9.

11 HEATING, VENTILATING, AND AIR CONDITIONING where 60 min/h 13.5 specific volume of moist air at 70 F db and 50% rh Equations for Steam Trap Selection The selection of the trap for the steam mains or risers is dependent on the pipe warm-up load and the radiation load from the pipe. Warm-up load is the condensate which is formed by heating the pipe surface when the steam is first turned on. For practical purposes, the final temperature of the pipe is the steam temperature. Warmup load is determined from W(tƒ t i)(0.114) C1 (74) ht where C 1 warm-up condensate, lb/h W total weight of pipe, lb (from tables in engineering handbooks) t ƒ final pipe temperature, F (steam temp.) t i initial pipe temperature, F (usually room temp.) specific heat constant for wrought iron or steel pipe (0.092 for copper tubing) h l latent heat of steam, Btu/lb (from steam tables) T time for warm-up, h The radiation load is the condensate formed by unavoidable radiation loss from a bare pipe. This load is determined from the following equation and is based on still air surrounding the steam main or riser: l LK(tƒ t i) C (75) where C 2 radiation condensate, lb/h L linear length of pipe, ft K heat transmission coefficient, Btu/(h lin ft F) 2 The radiation load builds up as the warm-up load drops off under normal operating conditions. The peak occurs at the midpoint of the warm-up cycle. Therefore, one-half of the radiation load is added to the warm-up load to determine the amount of condensate that the trap handles. h l Safety Factor Good design practice dictates the use of safety factors in steam trap selection. Safety factors from 2 to 1 to as high as 8 to 1 may be required, and for the following reasons: 1. The steam pressure at the trap inlet or the back-pressure at the trap discharge may vary. This changes the steam trap capacity. 2. If the trap is sized for normal operating load, condensate may back up into the steam lines or apparatus during start-up or warm-up operation.

12 16.12 ENVIRONMENTAL CONTROL 3. If the steam trap is selected to discharge a full and continuous stream of water, the air could not be vented from the system. The following guide is used to determine the safety factor: Design Safety factor Draining steam main 3 to 1 Draining steam riser 2 to 1 Between boiler and end of main 2 to 1 Before reducing valve 3 to 1 Before shutoff valve (closed part of time) 3 to 1 Draining coils 3 to 1 Draining apparatus 3 to 1 When the steam trap is to be used in a high-pressure system, determine whether the system is to operate under low-pressure conditions at certain intervals such as nighttime or weekends. If this condition is likely to occur, then an additional safety factor should be considered to account for the lower pressure drop available during nighttime operation. DETERMINING COOLING-TOWER FAN HORSEPOWER REQUIREMENTS A cooling tower serving an air-conditioning installation is designed for these conditions: Water flowrate, L 75,000 gal/min (4733 L/s); inlet water temperature, T i 110 F (43.3 C); outlet water temperature, T o 90 F (32.2 C); atmospheric wet-bulb temperature, T w 82 F (27.8 C); total fan efficiency as given by tower manufacturer, E T 75%; recirculation of air in tower, given by tower manufacturer, R c 8.5%; total air pressure drop through the tower, as given by manufacturer, P in (1.21 cm) H 2 O. What is the required fan horsepower input under these conditions? If the weather changes and the air outlet temperature becomes 102 F (38.9 C) with a wet-bulb temperature of 84 F (28.9 C)? Calculation Procedure: 1. Determine the fan horsepower for the given atmospheric and flow conditions The fan brake horsepower input for a cooling tower is given by the relation: BHP L T V sp R c P/ H 6356 E T, where T T o T i ; V sp specific volume of outlet air, ft 3 /lb (m 3 /kg); H difference between enthalpy of outlet air and inlet air, Btu/ lb (kj/ kg); other symbols as given earlier. Determine the enthalpy difference between the outlet air and inlet air by referring to a psychrometric chart where you will find that the enthalpy of the outlet air for the first case above, H o 72 Btu/lb (167.5 kj/kg); from the same source the enthalpy of the inlet air, H i 46 Btu/lb (107.0 kj/kg); likewise, V sp 15.1 ft 3 /lb (93.7 m 3 / kg) from the chart. Substituting in the equation above, BHP (75, )/(72 24) ; say 789 hp

13 HEATING, VENTILATING, AND AIR CONDITIONING (588.3 kw). In the above equation the constants and are used to convert gal/min to lb/min and air flow to ft 3 /lb, respectively. 2. Determine the power input required for the second set of conditions For the second set of conditions the air outlet temperature, T o 102 F (38.9 C) and the wet-bulb temperature is 84 F (28.9 C). Using the psychometric chart again, H Btu/lb (62.8 kj/kg), and the specific volume of the air at this temperature from the chart 15.2 ft 3 /lb (0.95 m 3 /kg). Substituting as before, BHP 764 hp (569.6 kw). Related Calculations. This procedure can be used with any type of cooling tower employed in air conditioning, steam power plants, internal combustion engines, or gas turbines. The method is based on knowing the tower s air outlet temperature. Use the psychometric chart to determine volumes and temperatures for various air states. As presented here, this method is the work of Ashfaq Noor, Dawood Hercules Chemicals Ltd., as reported in Chemical Engineering magazine. With the greater environmental interest in reducing stream pollution of all types, including thermal, cooling towers are receiving more attention as a viable way to eliminate thermal problems in streams and shore waters. The cooling tower is a nonpolluting device whose only environmental impact is the residue left in its bottom pans. Such residue is minor in amount and easily disposed of in an environmentally acceptable manner. CHOOSING AN ICE STORAGE SYSTEM FOR FACILITY COOLING Select an ice storage cooling system for a 100-ton (350-kW) peak cooling load, 10-h cooling day, 75 percent diversity factor, $8.00/ month kw demand charge, 12- month ratchet i.e., the utility term for a monthly electrical bill surcharge based on a previous month s higher peak demand. Analyze the costs for a partial-storage and for a full-storage system. Calculation Procedure: 1. Analyze partial-storage and full-storage alternatives Stored cooling systems use the term ton-h instead of tons of refrigeration, which is the popular usage for air-conditioning loads. Figure 1 shows a theoretical cooling load of 100 tons (350 kw) maintained for 100 h, or a 1000 ton-h (3500 kwh) cooling load. Each of the squares in the diagram represents 10 ton-h (35 kwh). No building air-conditioning system operates at 100 percent capacity for the entire daily cooling cycle. Air-conditioning loads peak in the afternoon, generally from 2:00 to 4:00 pm, when ambient temperatures are highest. Figure 2 shows a typical building air-conditioning load profile during a design day. As Fig. 2 shows, the full 100-ton chiller capacity (350 kw) is needed for only 2 of the 10 h in the cooling cycle. For the other 8 h, less than the total chiller capacity is required. Counting the tinted squares shows only 75, each representing 10 ton-h (35 kwh). The building, therefore, has a true cooling load of 750 ton-h (2625 kwh). A 100-ton (350 kw) chiller must however, be specified to handle the peak 100-ton (250 kw) cooling load.

14 16.14 ENVIRONMENTAL CONTROL Tons Hours FIGURE 1 Cooling load of 100 tons (351.7 kw) maintained for 10 h, or a 1000 ton-h cooling load. (Calmac Manufacturing Corporation.) Tons Hours FIGURE 2 Typical building air-conditioning load profile during a design day. (Calmac Manufacturing Corporation.) The diversity factor, defined as the ratio of the actual cooling load to the total potential chiller capacity, or diversity factor, percent 100 (Actual ton-hours)/total potential ton-hours. For this installation, diversity factor 100(750)/ percent. If a system s diversity factor is low, its cost efficiency is also low. Dividing the total ton-hours of the building by the number of hours the chiller is in operation gives the building s average load throughout the cooling period. If the air-conditioning load can be shifted to off-peak hours or leveled to the average load, 100 percent diversity can be achieved, and better cost efficiency obtained. When electrical rates call for complete load shifting, i.e., are excessively high, a conventionally sized chiller can be used with enough energy storage to shift the entire load into off-peak hours. This is called a full-storage system and is used most often in retrofit applications using existing chiller capacity. Figure 3 shows the same building air-conditioning load profile but with the cooling load completely shifted into 14 off-peak hours. The chiller is used to build and store ice during the night. The 32 F (0 C) energy stored in the ice then provides the required 750 ton-h (2625

15 HEATING, VENTILATING, AND AIR CONDITIONING Tons Hours FIGURE 3 Building air-conditioning load profile of Fig. 2 with the cooling load shifted into 14 off-peak hours. (Calmac Manufacturing Corporation.) kwh) of cooling during the day. The average load is lowered to (750 ton-h)/14 h 53.6 tons (187.6 kw), which results in significantly reduced demand charges. In new construction, a partial-storage system is the most practical and cost effective load-management strategy. In this load-leveling method, the chiller runs continuously. It charges the ice storage at night and cools the load directly during the day with help from stored cooling. Extending the hours of operation from 14 to 24 h results in the lowest possible Average Load, (750 ton-h)/24 hours tons (109.4 kw, as shown by the plot in Fig. 4). Demand charges are greatly reduced and chiller capacity can often be decreased by 50 to 60 percent, or more. 2. Compute partial-storage demand savings Cost estimates for a conventional chilled-water air-conditioning system comprised of a 100-ton (350 kw) chiller with all accessories such as cooling tower, fan coils, Tons Hours FIGURE 4 Extending the hours of operation for 14 to 24 results in the lowest possible average load 750 ton-h/ Demand charges are greatly reduced and chiller capacity can often be decreased 50 to 60 percent, or more. (Calmac Manufacturing Corporation.)

16 16.16 ENVIRONMENTAL CONTROL pumps, blowers, piping, controls, etc., show a price of $600/ton, or 100 tons $600/ton $60,000. The distribution system for this 100-ton (350-kW) plant will cost about the same, or $60,000. Total cost therefore $60,000 $60,000 $120,000. With partial-storage using a 40 percent size chiller with ice storage at a 75 percent diversity factor, the true cooling load translates into 750 ton-h (2626 kwh) with the chiller providing 400 ton-h (1400-kWh) and stored ice the balance, or ton-h (1225 kwh). Hence, cost of 40-ton chiller at $600/ton $24,000. From the manufacturer of the stored cooling unit, the installed cost is estimated to be $60/ton-h, or $ ton-h $21,000. The distribution system, as before, costs $60,000. Hence, the total cost of the partial-storage system will be $24,000 $21,000 $60,000 $105,000. Therefore, the purchase savings of the partial-storage system over the conventional chilled-water air-conditioning system $120,000 $105,000 $15,000. The electrical demand savings, which continue for the life of the installation, are: (100 tons 40 tons chiller capacity)(1.5-kw/ton at peak summer demand conditions, including all accessories)($8.00/mo/kw demand charge)(12 mo/yr) $ Determine full-storage savings With full-storage, 100-ton (350-kW) peak cooling load, 10-h cooling day, 75 percent diversity factor, 1000-h cooling season, $8.00/ mo/ kw demand charge, 12- month ratchet, $0.03/ kwh off-peak differential, the chiller cost will be (10 h)(100 tons)(75 percent)($60/ton-h, installed) $45,000. The demand savings will be, as before, (100 tons)(1.5 kw/ton)(12 mo)($8.00/mo/kw demand charge)($8.00) $14,400/ year. Energy savings are computed using the electric company s off-peak kwh off-peak differential, or ($0.03/ kwh)(1000 h)(100 tons)(1.2 average kw/ ton) $3,600/year. The simple payback time for this project (equipment cost, $)/ (demand savings, $ energy savings, $) $45,000/$18, yr. After the end of the payback time there is an annual energy savings of $18,000/year. And as rates increase, which they usually do, the annual savings will probably increase above this amount. Related Calculations. Ice storage systems are becoming more popular for a variety of structures: office buildings, computer data centers, churches, nursing homes, police stations, public libraries, theaters, banks, medical centers, hospitals, hotels, convention centers, schools, colleges, universities, industrial training centers, cathedrals, medical clinics, manufacturing plants, warehouses, museums, country clubs, stock exchanges, government buildings, and courthouses. There are several reasons for this growing popularity: (1) utility power costs can be reduced by shifting electric power demand to off hours by avoiding peak-demand charges; (2) lower overall electric rates can be obtained for the facility if the kilowatt demand is reduced, thereby eliminating the need for the local utility to build new generating facilities; (3) ice storage can provide uninterrupted cooling in times of loss of outside, or inside, electric generating capability during natural disasters, storms, or line failures the ice storage system acts like a battery, giving the cooling required until the regular coolant supply can be reactivated; (4) environmental regulations are more readily met because less power input is required, reducing the total energy usage; (5) by making ice at night, the chiller operates when the facilities electrical demands are lowest and when a utility s generating capacity is underutilized; (6) provision can be made to use more environmentally friendly HCFC- 123, thereby complying with current regulations of federal and state agencies; (7) facility design can be planned to include better control of indoor air quality, another

17 HEATING, VENTILATING, AND AIR CONDITIONING environmentally challenging task faced by designers today; (8) new regulations, specifically The Energy Policy Act of 1992 (EPACT), curtails the use of and eliminates certain fluorescent and incandescent lamps (40-W T12, cool white, warm white, daylight white, and warm white deluxe), which will change both electrical demand and replacement bulb costs in facilities, making cooling costs more important in total operating charges. Designers now talk of greening a building or facility, i.e., making it more environmentally acceptable to regulators and owners. An ice storage system is one positive step to greening a facility while reducing the investment required for cooling equipment. The procedure given here clearly shows the savings possible with a typical well-designed ice storage system. There are three common designs used for ice storage systems today: (1) directexpansion ice storage where ice is frozen directly on metal refrigerant tubes submerged in a water tank; cooled water in the tank is pumped to the cooling load when needed; (2) ice harvester system where a thin coat of ice is frozen on refrigerated metal plates and periodically harvested into a bin or water tank by melting the bond of the ice to the metal plates; the chilled water surrounding the ice is pumped to the cooling load when needed; (3) patented ice bank system uses a modular, insulated polyethylene tank containing a spiral-wound plastic tube heat exchanger surrounded with water; at night a 26 F ( 3.33 C) 75 percent water/25 percent glycol solution from a standard packaged air-conditioning chiller circulates through the heat exchanger, freezing solid all the water in the tank; during the day the ice cools the solution to 44 F (6.66 C) for use in the air-cooling coils where it cools the air from 75 F (23.9 C) to 55 F (12.8 C). The patented system has several advantages over the first two, namely: (1) ice is the storage medium, rather than water. One pound (0.45 kg) of ice can store 144 Btu (152 J) of energy, while one pound of water in a stratified tank stores only 12 to 15 Btu (12.7 J to 15.8 J). Hence, such an ice storage system needs only about one-tenth the space for energy storage. This small space requirement is important in retrofit applications where space is often scarce. (2) Patented systems are closed; there is no need for water treatment or filtration; pumping power requirements are small; (3) power requirements are minimal; (4) installation of the insulated modular tanks is fast and inexpensive since there are no moving parts; the tanks can be installed indoors or outdoors, stacked or buried to save space. Currently these tanks are available in three sizes: 115, 190, and 570 ton-h (402.5, 665, and 1995 kwh). (5) A low-temperature duct system can be used with 45 F (7.2 C) air instead of the conventional 55 F (12.8 C) air in the airconditioning system. This can permit further large savings in initial and operating costs. The 45 F (7.2 C) primary air requires much lower air flow [ft 3 /min (m 3 /m)] than 55 F (12.8 C) air. This reduces the needed size of both the air handler and duct system; both may be halved. Energy savings from the smaller air-handler motors may total 20 percent, even after figuring the additional energy required for the small mixing-box motors. This procedure is based on data provided by the Calmac Manufacturing Corporation, Englewood, NJ. The economic analysis was provided by Calmac, as were the illustrations in this procedure. Calmac manufactures the Levload modular insulated storage tanks mentioned above that are used in their Ice Bank Stored Cooling System. Their system, when designed with a low-temperature heat-recovery loop, can also make the chiller into a water-source heat pump for winter heating. Thus, office and similar buildings often require heating warm-up in the morning on winter days, but these same buildings may likewise require cooling in the afternoon because of lights, people, computers, etc. Ice made in the morning to provide heat-

18 16.18 ENVIRONMENTAL CONTROL ing supplies free afternoon cooling and melts to be ready for the next day s warmup. Even on coldest days, low-temperature waste heat (such as cooling water or exhaust air), or off-peak electric heat can be used to melt the ice. Oil or gas connections to the building can thus be eliminated. Nontoxic eutectic salts are available to lower the freezing point of the water in Calmac Ice Banks to either 28 F ( 2.2 C) or 12 F ( 11.1 C) and, consequently, the temperature of the resulting ice. Twenty-eight-degree ice, for example, can provide cold, dry primary air for many uses, including extra-low temperature airside applications. Twelve-degree ice can be used for on-ground aircraft cooling, off-peak freezing of ice rinks, and for industrial process applications requiring colder liquids. Other temperatures can be provided for specialized applications, such as refrigerated warehouses. Figure 5 shows the charge cycle using a partial storage system for an airconditioning installation. At night a water-glycol solution is circulated through a standard packaged air-conditioning chiller and the Ice Bank heat exchanger, bypassing the air-handler coil. The cooling fluid is at 26 F ( 3.3 C) and freezes the water surrounding the heat exchanger. During the day, Fig. 6, the water-glycol solution is cooled by the Ice Bank from 52 F (11.1 C) to 34 F (1.1 C). A temperature-modulating valve, set at 44 F (6.7 C) in a bypass loop around the Ice Bank, allows a sufficient quantity of 52 F (11.1 C) fluid to bypass the Ice Bank, mix with 34 F (1.1 C) fluid, and achieve the desired 44 F (6.7 C) temperature. The 44 F (6.7 C) fluid enters the coil, where it cools the air passing over the coil from 75 F (23.9 C) to 55 F (12.8 C). Fluid leaves the coil at 60 F (15.6 C), enters the chiller and is cooled to 52 F (11.1 C). Note that, while making ice at night, the chiller must cool the water-glycol to 26 F ( 3.3 C), rather than produce 44 F (6.7 C) or 45 F (7.2 C) water temperatures required for conventional air-conditioning systems. This has the effect of derating the nominal chiller capacity by about 30 percent. Compressor efficiency, however, is only slightly reduced because lower nighttime temperatures result in cooler condenser water from the cooling tower (if used) and help keep the unit operating efficiently. Similarly, air-cooled chillers benefit from cooler condenser entering airtemperatures at night. The temperature-modulating valve in the bypass loop has the added advantage of providing unlimited capacity control. During many mild-temperature days in the FIGURE 5 Counterflow heatexchanger tubes used in the Ice Bank. (Calmac Manufacturing Corporation.)

19 HEATING, VENTILATING, AND AIR CONDITIONING SI Values F C Charge cycle. (Calmac Manufacturing Cor- FIGURE 6 poration.) spring and fall, the chiller will be capable of providing all the cooling needed for the building without assistance from stored cooling. When the building s actual cooling load is equal to or lower than the chiller capacity, all of the system coolant flows through the bypass loop, as in Fig. 8. Using 45 F (7.2 C) rather than 55 F (12.8 C) system air in the air-conditioning system permits further large savings in initial and operating costs. The 45 F (7.2 C) SI Values F C Discharge cycle. (Calmac Manufacturing Cor- FIGURE 7 poration.)

20 16.20 ENVIRONMENTAL CONTROL SI Values F C FIGURE 8 When cooling load equals, or is lower than chiller capacity, all coolant flows through bypass loop. (Calmac Manufacturing Corporation.) low-temperature air is achieved by piping 38 F (3.3 C) water-glycol solution from the stored cooling Ice Bank to the air handler coil instead of mixing it with bypassed solution, as in Fig. 6. The 45 F (7.2 C) air is used as primary air and is distributed to motorized fan-powered mixing boxes where it is blended with room air to obtain the desired room temperature. Primary 45 F (7.2 C) air requires much lower ft 3 / min (m 3 /min) than 55 F (12.8 C) air. Consequently, the size and cost of the air handler and duct system may be cut in about half. Energy savings of the smaller air handler motors total 20 percent, even counting the additional energy required for the small mixing-box motors. The recommended coolant solution for these installations is an ethylene glycolbased industrial coolant such as Union Carbide Corporation s UCARTHERM or Dow Chemical Company s DOWTHERM SR-1. Both are specially formulated for low viscosity and superior heat-transfer properties, and both contain a multicomponent corrosion inhibitor system effective with most materials of construction, including aluminum, copper, solder, and plastics. Standard system pumps, seals, and air-handler coils can be sued with these coolants. However, because of the slight difference in the heat-transfer coefficient between water-glycol and plain water, air-handler coil capacity should be increased by about 5 percent. Further, the water and glycol must be thoroughly mixed before the solution enters the system. Another advantage of ice storage systems for cooling and heating is provision of an uninterrupted power supply (UPS) in the event of the loss of a building s cooling or heating facilities. Such an UPS can be important in data centers, hospitals, research laboratories, and other installations where cooling or heating are critical. Figure 9a shows the conventional ice builder and Fig. 9b the LEVLOAD Ice Bank. When ice is stored remote from the refrigerating system evaporator, as in Fig. 9b, the evaporator is left free to aid the cooling during the occupied hours of a building or other structure. Figure 10 compares the chiller performance of a Partial Storage Ice Bank, Fig. 9b (upper curve), with an ice-builder system, Fig. 9a (lower curve), on a typical design day. Note that when compressor cooling is done through ice on the evaporator, suction temperatures are low and kw/ ton is increased.

21 HEATING, VENTILATING, AND AIR CONDITIONING Ice Builder Solid Ice Melted Ice Area System Coolant (Water-Glycol) Refrigerant System Coolant (Water) LEVLOAD Ice Bank FIGURE 9 (a) Typical ice-builder arrangement. (b) LEVLOAD Ice Bank method of ice burn-off (ice melting). (Calmac Manufacturing Corporation.) SI Values F C FIGURE 10 Comparison of chiller performance of a Partial Storage Ice Bank (upper curve) and conventional ice-builder system (lower curve). (Calmac Manufacturing Corporation.) With discussions still taking place about chlorofluorocarbon refrigerants suitable for environmental compliance, designers have to seek the best choice for the system chiller. One approach finding popularity today as an interim solution is to choose a chiller which can use an energy-efficient refrigerant today and the most environmentally friendly refrigerant in the future. Thus, for some chillers, CFC-11 is the most energy-efficient refrigerant today. The future most environmentally friendly refrigerant is currently thought to be HCFC-123. By choosing and sizing a chiller that can run on HCFC-123 in the future, energy savings can be obtained today,

22 16.22 ENVIRONMENTAL CONTROL and, if environmental requirements deem a switch in the future, the same chiller can be used for CFC-11 today and HCFC-123 in the future. It is also possible that the same chiller can be retrofitted to use a non-cfc refrigerant in the future. A number of ice storage systems have adopted this design strategy. Many firms that installed ice storage systems in recent years are so pleased with the cost savings (energy, equipment, ducting, UPS, etc.) that they plan to expand such systems in the future. Chillers using CFC-11 normally produce ice at 0.64 to 0.75 kw/ ton, depending on the amount of ice produced. Power consumption is lower when larger quantities of ice are produced. When HCHC-123 is used, the power input ranges between 0.7 and 0.8 kw/ton, again depending on the number of tons produced. As before, power consumption is lower when larger tonnages are produced. New centrifugal chillers produce cooling at power input ranges close to 0.5 kw/ ton. In all air-conditioning systems, designers must recognize that there are three courses of action open to them when CFC refrigerants are no longer available: (1) continue to use existing CFC-based equipment, taking every precaution possible to stop leaks and conserve available CFC supplies; (2) retrofit existing chilling equipment to use non-cfc refrigerants; this step requires added investment and changed operating procedures; (3) replace existing chillers with new chillers specifically designed for non-cfc refrigerants; again, added investment and changed operating procedures will be necessary. To avoid CFC problems, new high-efficiency chlorine-free screw chillers are being used. And there are packaged ammonia screw chiller available also. Likewise, a variety of alternative refrigerants are now being produced for new and retrofit refrigeration and air-conditioning uses. Table 1 shows an economic analysis of typical partial-storage and full-storage installations. A conventional chilled-water air-conditioning system is compared with a partial-storage 40 percent size chiller with Ice Banks in the partial-storage analysis. Full-storage produces the simple payback time of 2.5 years for the investment. Data in this analyze are from Calmac Manufacturing Corporation. When using a similar analysis, be certain to obtain current prices for components, demand charges, and electricity. Values given here are for illustration purposes only. ANNUAL HEATING AND COOLING ENERGY LOADS AND COSTS A 2000 ft 2 (185.8 m 2 ) building has a 100,000 Btu/h (29.3 kw) heat loss in an area where the heating season is 264 days duration. Average winter outdoor temperature is 42 F (5.6 C); design conditions are 70 F (21.1 C) indoors and 0 F ( 17.8 C) outdoors. The building also has a summer cooling load of 7.5 tons (26.4 kw) with an estimated full-load cooling time of 800 operating hours. What are the total winter and summer estimated loads in Btu/h (kw)? If oil is 90 cents/gallon and electricity is 7 cents/kwh, what are the winter and summer energy costs? Use a 24-h heating day for winter loads and a boiler efficiency of 75 percent when burning oil with a higher heating value (HHV) of 140,000 Btu/ gal (39,018 MJ/ L). Calculation Procedure: 1. Compute the winter operating costs The winter seasonal heating load, WL N 24 h/day Btu/h heat loss (average indoor temperature, F average outdoor temperature, F)/(average in-

23 TABLE 1 Economic Analysis of Typical Partial Storage and Full Storage Installations* Partial storage Assume: 100-ton peak cooling load, 10-h cooling day, 75 percent diversity factor, $8.00/mo/kW demand charge, 12-mo ratchet.* Conventional Chilled Water Air Conditioning System: 100-ton chiller at $600/ ton, installed** $ 60,000 Distribution system 60,000 Total $120,000 Partial storage (40% size chiller with Ice Banks): At 75 percent diversity factor, the true cooling load translates into 750 tonh with the chiller providing 400 ton-h and stored cooling the balance, or 350 ton-h. Therefore: 40-ton chiller at $600/ ton, installed $ 24,000 Stored cooling at $60/ ton-h, installed 21,000 Distribution system 60,000 Total $105,000 Purchase savings: $ 15,000 Demand savings: 60 tons 1.5 kw/ton 12 mo $8.00 $8,640/yr *Utility term for a monthly electrical bill surcharge based on a previous month s higher peak demand. Model Model Model Specifications Total ton-hour capacity Tube surface/ton-h, ft Nominal discharge time, h Latent storage cap., ton-h Sensible storage cap., ton-h Maximum operating temp., F Maximum operating press., lb/ Outside diameter, in Length width, in Height, in Weight, unfilled, lb 1,060 1,550 4,850 Weight, filled, lb ,750 50,450 Full storage Assume: 100-ton peak cooling load, 10-hour cooling day, 75% diversity factor, 1000-hour cooling season, $8.00/mo/kW demand charge, 12-mo ratchet, $0.03/ kwh off-peak differential. Full storage: 10 h 100 tons 75% $60/ton-h, installed $45,000 Demand savings: 100 tons 1.5 kw/ton 12 mo $8.00 $14,400/yr Energy savings: $0.03/kWh 1000 h 100 tons 1.2 Avg. kw/ton $ 3600/yr Total savings: $18,000/ yr Simple payback: $45,000 $18, yr. **The $600/ ton includes all accessories, such as cooling tower, fan coils, pumps, blowers, piping, controls, etc. Figure shown is for conventional temperature system. This cost could be reduced by 50 percent by using a low temperature duct system. The 1.5 kw/ton is figured at the peak summer demand conditions and also includes all accessories. Model Model Model Specifications Volume of water/ ice, gals Volume of solution in HX, gals Press. drop (25% glycol, 28 F), PSI 20 gal/min gal/ min gal/ min gal/ min gal/min gal/ min 13.0 The outlet temperatures from the tanks vary with the rate at which the tanks are discharged. See LEVLOAD Performance Manual for details. LEVLOAD and CALMAC are registered trademarks of Calmac Manufacturing Corporation. The described product and its applications are protected by United States Patents 4,294,078; 4,403,645; 4,565,069; 4,608,836; 4,616,390; 4,671,347 and 4,687,588. *Calmac Manufacturing Corporation SI values in procedure text

24 16.24 ENVIRONMENTAL CONTROL door temperature, F outside design temperature, F), where N number of days in the heating season. For this building, WL ( ,000)(70 42)/ (70 0) 253,440,000 Btu (267.4 MJ). The cost of the heating oil, CO Btu seasonal heating load oil cost $/gal/ boiler efficiency HHV. Or, for this building, CO (253,440,000)(0.90)/ ,000 $2, for the winter heating season. 2. Calculate the summer cooling cost The summer seasonal electric consumption in kwh, SC tons of air conditioning kw/ton of air conditioning number of operating hours of the system. The kw/ design ton factor is based on both judgment and experience. General consensus amongst engineers is that the kw/ ton varies from 1.8 for small window-type systems to 1.0 for large central-plant systems. The average value of 1.4 kw/ design ton is frequently used and will be used here. Thus, the summer electric consumption, SC kwh. This energy will cost 8400 kwh $0.07 $ The total annual energy cost is the sum of the winter and summer energy costs, $ $ $ Related Calculations. Use this procedure to compute the energy costs for any type of structure, industrial, office, residential, medical, educational, etc., having heating or cooling loads, or both. Any type of fuel, oil, gas, coal, etc., can be used for the structure. This procedure is the work of Jerome F. Mueller, P.E. of Mueller Engineering Corp. HEAT RECOVERY USING A RUN-AROUND SYSTEM OF ENERGY TRANSFER A hospital operating room suite requires 6000 ft 3 /min (169.8 m 3 /s) of air in the supply system with 100 percent exhaust and 100 percent compensating makeup air. Winter outdoor design temperature is 0 F ( 17.8 C); operating-room temperature is 80 F (26.7 C) with 50 percent relative humidity year-round. How much energy can be saved by installing coils in both the supply and exhaust air ducts with a pump circulating a non-freeze liquid between the two coils, absorbing heat from the exhaust air and transferring this heat to the makeup air being introduced? Calculation Procedure: 1. Choose the coils to use In the winter the exhaust air is at 80 F (26.7 C) while the supply air is at 0 F ( 17.8 C). Hence, the coil in the exhaust air duct will transfer heat to the nonfreeze liquid. When this liquid is pumped through the coil in the intake-air duct it will release heat to the incoming air. This transfer of otherwise wasted heat will reduce the energy requirements of the system in the winter. As a first choice, select a coil area of 12 ft 2 (1.1 m 2 ) with a flow of 6000 ft 3 / min (169.8 m 3 /s). While a number of coil arrangements are possible, the listing below shows typical coil conditions at face velocities of 500 ft/min (152.4 m/min) and 600 ft/min (182.8 m/min) with a coil having 8-fins/in coil. Entering coolant

25 HEATING, VENTILATING, AND AIR CONDITIONING temperature 45 F (7.2 C); entering air temperature 80 F (26.7 C) dry bulb, 67 F (19.4 C) wet bulb. Various manufacturers values may vary slightly from these values. Ft/min (m/min) face velocity Temp rise, F ( C) No. of rows Total MBtu/h (kwh) Leaving dry bulb, F ( C) Leaving wet bulb, F ( C) 500 (152.4) 12 (21.6) (4.86) 57.3 (14.1) 56.5 (13.6) 500 (152.4) 12 (21.6) (5.89) 54.2 (12.3) 53.9 (12.2) 600 (182.9) 10 (18) (4.19) 62.1 (16.7) 59.7 (15.4) The middle coil listed above, if placed in the exhaust duct, would produce 20.1 MBtu/h (5.89 kwh) with a 12 F (21.6 C) temperature rise with a leaving air temperature of 53.9 F (12.2 C) when the liquid coolant enters the coil at 45 F (7.2 C) and leaves 57 F (13.9 C). 2. Determine the coil heating capacity The heating capacity of a coil is the product of (coil face area, ft 2 )[heat release, Btu/(h ft 2 ) of coil face area]. For this coil, heating capacity 12 20, ,200 Btu/h (70.7 kwh). The incoming makeup air can be heated to a temperature of: (heating capacity, Btu)/(makeup air flow, ft 3 /min)(1.08) 241,200/ F (2.9 C). Hence, the makeup air is heated from 0 F ( 17.8 C) to 37.2 F (2.9 C). The energy saved is assuming 1000 Btu/ lb of steam (2330 kj/ kg) 241,200 / lb/h (109.5 kg/h). With a 200-day heating season and 10-h operation/day, the saving will be ,400 lb/yr (219,010 kg/yr). And if steam costs $20/thousand pounds ($20/454 kg), the saving will be (482,400/1000)($20) $9,648,00/year. Such a saving could easily pay for the heating coil in one year. Related Calculations. Use this general approach to choose heating coils for any air-heating application where waste heat can be utilized to increase the temperature of incoming air, thereby reducing the amount of another heating medium that might be required. Most engineers use the 1000 Btu/lb (2300 kj/kg) latent heat of steam as a safe number to convert quickly from hourly heat savings in Btu (kg) to pounds of steam. This procedure can be used for industrial, commercial, residential, and marine applications. The procedure is the work of Jerome F. Mueller, P.E., of Mueller Engineering Corp. Figure 11 is a typical run-around coil detail that is very commonly used in energy recovery systems in which the purpose is to exrract heat from air that must be exhausted. Normally about 40 to 60 percent of the heat being wasted can be recovered. This seemingly simple detail has two points that should be carefully noted. The difference in fluid temperatures in this closed system creates small expansion and contraction problems and an expansion tank is required. Most importantly the temperature of the incoming supply air can create a coil temperature so low that the coil in the exhaust air stream begins to ice up. Normally this betins at some 35 F (1.67 C). This is when the three-way bypass valve comes into play and the glycol is not circulated through the outside air coil. Obviously if the outside air temperature is low enough, the system will go into full bypass and the ciruclating pump should be stopped.