1999 HVAC Applications Part 10 ppsx

1.10 1999 ASHRAE Applications Handbook (SI) Capital and interest Salvage value Replacements Operating energy Property tax Maintenance Insurance Interest deduction Table 6 summarizes the interest and principle payments for this example. Annual payments are the product of the initial system cost C s,init and the capital recovery factor CRF(i m ,5). Also, Equation (10) can be used to calculate total discounted interest deduction directly. Next, apply the capital recovery factor CRF(i ′,5) and tax rate T inc to the total of the discounted interest sum. Depreciation Use the straight line depreciation method to calculate depreciation: Next, discount the depreciation. Finally, the capital recovery factor and tax are applied. U.S. tax code recommends estimating the salvage value prior to depreciating. Then depreciation is claimed as the difference between the initial and salvage value, which is the way depreciation is treated in this example. The more common practice is to initially claim zero salvage value, and at the end of ownership of the item, treat any salvage value as a capital gain. C sinit, ITC– () CRF i ′ n (,) $10 000 $0– () 0.229457 $2294.57== C ssalv, PWF i ′ n (,)CRF i ′ n (,)1 T salv – () $1000 0.792471 × 0.229457 × 0.5 × $90.92== R k PWF i ′ k (,)[] CRF i ′ n (,)1 T inc – () k 1= n ∑ $500 0.869741 × 0.229457 × 0.5 × $49.89== C e CRF i ′ n (,)CRF i ″ n (,)⁄[] 1 T inc – () $500 0.229457 0.211247 ⁄[] 0.5 $271.55== C sassess, T prop 1 T inc – () $10 000 0.40 × 0.01 × 0.5 × $20.00== M 1 T inc – () $100 1 0.5– () $50.00== I 1 T inc – () $50 1 0.5– () $25.00== T inc i m P k 1– PWF i d k (,)[] k 1= n ∑ CRF i ′ n (,) … see Table 6= Year D k,SL PWF(i d ,k) Discounted Depreciation 1 $1800.00 0.909091 $1636.36 2 $1800.00 0.826446 $1487.60 3 $1800.00 0.751315 $1352.37 4 $1800.00 0.683013 $1229.42 5 $1800.00 0.620921 $1117.66 Total $6823.42 $2554.66 CRF i ′ 5(,)T inc $2554.66 0.229457 × 0.5 × $293.09== T inc D kSL, PWF i d k (,)[] CRF i ′ n (,)… k 1= n ∑ D kSL, C sinit, C ssalv, – () n ⁄ $10 000 $1000– () 5 ⁄ $1800.00== = $6823.42 CRF i ′ n (,) T inc $6823.42 0.229457 × 0.5 × $782.84== Table 6 Interest Deduction Summary (for Example 9) Year Payment Amount, Current $ Interest Payment, Current $ Principal Payment, Current $ Outstanding Principal, Current $ PWF(i d , k) Discounted Interest, Discounted $ Discounted Payment, Discounted $ 0 — — — 10 000.00 — — — 1 2 637.97 1 000.00 1 637.97 8 362.03 0.909091 909.09 2 398.17 2 2 637.97 836.20 1 801.77 6 560.26 0.826446 691.07 2 180.14 3 2 637.97 656.03 1 981.95 4 578.31 0.751315 492.89 1 981.95 4 2 637.97 457.83 2 180.14 2 398.17 0.683013 312.70 1 801.77 5 2 637.97 239.82 2 398.17 0 0.620921 148.91 1 637.97 _________________ ___________________ _________________ ____________________ Total — 3 189.88 10 000.00 — — 2 554.66 10 000.00 Table 7 Summary of Cash Flow (for Example 10) 1234567 89 10 11 Yea r Cash Outlay, $ Net Income Before Taxes, $ Depreciation, $ Net Taxable Income, a $ Income Taxes @50%, $ Net Cash Flow, b $ Present Worth of Net Cash Flow 10% Rate 15% Rate 20% Rate PWF P, $ P, $ P, $ 01200000000 −120 000 1.000 −120 000 −120 000 −120 000 1 0 20 000 15 000 5 000 2 500 17 500 0.909 15 900 15 200 14 600 2 0 30 000 15 000 15 000 7 500 22 500 0.826 18 600 17 000 15 600 3 0 40 000 15 000 25 000 12 500 27 500 0.751 20 600 18 100 15 900 4 0 50 000 15 000 35 000 17 500 32 500 0.683 22 200 18 600 15 700 5 0 50 000 15 000 35 000 17 500 32 500 0.621 20 200 16 200 13 100 6 0 50 000 15 000 35 000 17 500 32 500 0.564 18 300 14 100 10 900 7 0 50 000 15 000 35 000 17 500 32 500 0.513 16 700 12 200 9 100 8 0 50 000 15 000 35 000 17 500 32 500 0.467 15 200 10 600 7 600 Total Cash Flow 27 700 2 000 − 17 500 Investment Value 147 500 122 000 102 500 a Net taxable income = net income − depreciation. b Net cash flow = net income − taxes. Owning and Operating Costs 1.11 Summary of terms Cash Flow Analysis Method. The cash flow analysis method accounts for costs and revenues on a period-by-period (e.g., year- by-year) basis, both actual and discounted to present value. This method is especially useful for identifying periods when net cash flow will be negative due to intermittent large expenses. Example 10. An eight-year study for a $120 000 investment with depreciation spread equally over the assigned period. The benefits or incomes are variable. The marginal tax rate is 50%. The rate of return on the investment is required. Table 7 has columns showing year, cash outlays, income, depreciation, net taxable income, taxes and net cash flow. Solution: To evaluate the effect of interest and time, the net cash flow must be multiplied by the single payment present worth factor. An arbi- trary interest rate of 10% has been selected and the PWF sgl is obtained by using Equation (4). Its value is listed in Table 7, column 8. Present worth of the net cash flow is obtained by multiplying columns 7 and 8. Column 9 is then added to obtain the total cash flow. If year 0 is ignored, an investment value is obtained for a 10% required rate of return. The same procedure is used for 15% interest (column 10, but the PWF is not shown) and for 20% interest (column 11). Discussion. The interest at which the summation of present worth of net cash flow is zero gives the rate of return. In this example, the investment has a rate of return by interpolation of about 15.4%. If this rate offers an acceptable rate of return to the investor, the proposal should be approved; otherwise, it should be rejected. Another approach would be to obtain an investment value at a given rate of return. This is accomplished by adding the present worth of the net cash flows, but not including the investment cost. In the example, under the 10% given rate of return, $147 700 is obtained as an investment value. This amount, when using money that costs 10%, would be the acceptable value of the investment. Computer Analysis Many computer programs are available that incorporate the economic analysis methods described above. These range from simple macros developed for popular spreadsheet applications to more comprehensive, menu-driven computer programs. Commonly used examples of the latter include Building Life-Cycle Cost (BLCC), Life Cycle Cost in Design (LCCID), and PC-ECONPACK. BLCC was developed by the National Institute of Standards and Technology (NIST) for the U.S. Department of Energy (DOE). The program follows criteria established by the Federal Energy Management Program (FEMP) and the Office of Manage- ment and Budget (OMB). It is intended for the evaluation of energy conservation investments in nonmilitary government buildings; however, it is also appropriate for similar evaluations of commercial facilities. LCCID is an economic analysis program tailored to the needs of the U.S. Department of Defense (DOD). Developed by the U.S. Army Corps of Engineers and the Construction Engineering Research Laboratory (USA-CERL), LCCID uses economic criteria established by FEMP and OMB. PC-Econpack, developed by the U.S. Army Corps of Engineers for use by the DOD, uses economic criteria established by the OMB. The program performs standardized life-cycle cost calculations such as net present value, equivalent uniform annual cost, SIR, and discounted payback period. Macros developed for common spreadsheet programs generally contain preprogrammed functions for the various life-cycle cost calculations. Although typically not as sophisticated as the menu- driven programs, the macros are easy to install and easy to learn. Reference Equations Table 8 lists commonly used discount formulas as addressed by NIST. Refer to NIST Handbook 135 (Ruegg) and Table 2.3 in that handbook for detailed discussions. SYMBOLS c = cooling system adjustment factor C = total annual building HVAC maintenance cost C e = annual operating cost for energy C s,assess = assessed system value C s,init = initial system cost C s,salv = system salvage value at end of study period C y = uniform annualized mechanical system owning, operating, and maintenance costs CRF = capital recovery factor CRF(i,n) = capital recovery factor for interest rate i and analysis period n CRF(i ′,n) = capital recovery factory for interest rate i′ for items other than fuel and analysis period n CRF(i ″,n) = capital recovery factor for fuel interest rate i″ and analysis period n CRF(i m ,n) = capital recovery factor for loan or mortgage rate i m and analysis period n d = distribution system adjustment factor D k = depreciation during period k Capital and interest −$2294.57 Salvage value +$ 90.92 Replacements −$ 49.89 Operating costs −$ 271.55 Property tax −$ 20.00 Maintenance −$ 50.00 Insurance −$ 25.00 Interest deduction +$ 293.09 Depreciation deduction +$ 782.84 Total annualized cost −$1544.16 Table 8 Commonly Used Discount Formulas Name Algebrac Form a,b Single compound-amount (SCA) equation Single present value (SPW) equation Uniform sinking-fund (USF) equation Uniform capital-recovery (UCR) equation Uniform compound-account (UCA) equation Uniform present-value (UPW) equation Modified uniform present-value (UPW*) equation where A = end-of-period payment (or receipt) in a uniform series of payments (or receipts) over n periods at d interest or discount rate A 0 = initial value of a periodic payment (receipt) evaluated at the begin- ning of the study period A t = A 0 ·(1 + e) t , where t = 1,…, n d = interest or discount rate e = price escalation rate per period Source: NIST Handbook 135 (Ruegg). a Note that the USF, UCR, UCA, and UPW equations yield undefined answers when d= 0. The correct algebraic forms for this special case would be as follows: USF formula, A = F/N; UCR formula, A = P/N; UCA formula, F = A·n. The UPW* equation also yields an undefined answer when e = d. In this case, P = A 0 ·n. b The terms by which the known values are multiplied in these equations are the formulas for the factors found in discount factor tables. Using acronyms to represent the factor formulas, the discounting equaitons can also be written as F = P·SCA, P=F·SPW, A = F·USF, A = P·UCR, F = UCA, P = A·UPW, and P = A 0 ·UPW*. FP1 d+ () n []⋅ = PF 1 1 d+ () n ⋅ = AF d 1 d+ () n 1– ⋅ = AP d 1 d+ () n 1 d+ () n 1– ⋅ = FA 1 d+ () n 1– d ⋅ = PA 1 d+ () n 1– d 1 d+ () n ⋅ = PA 0 1 e+ de–   1 1 e+ 1 d+   n – ⋅⋅ = 1.12 1999 ASHRAE Applications Handbook (SI) D k,SL = depreciation during period k due to straight line depreciation method D k,SD = depreciation during period k due to sum-of-digits depreciation method F = future value of a sum of money h = heating system adjustment factor i = compound interest rate per period i d = discount rate per period i m = market mortgage rate i ′ = effective interest rate for all but fuel i ″ = effective interest rate for fuel I = insurance cost per period ITC = investment tax credit j = inflation rate per period j e = fuel inflation rate per period k = end of period(s) during which replacement(s), repair(s), depreciation, or interest are calculated M = maintenance cost per period n = number of periods under analysis P = present value of a sum of money P k = outstanding principle on loan at end of period k PMT = future equal payments PWF = present worth factor PWF(i d ,k) = present worth factor for discount rate i d at end of period k PWF(i ′,k) = present worth factor for effective interest rate i′ at end of period k PWF(i,n) sgl = single payment present worth factor PWF(i,n) ser = present worth factor for a series of future equal payments R k = net replacement, repair, or disposal costs at end of period k T inc = net income tax rate T prop = property tax rate T salv = tax rate applicable to salvage value of system REFERENCES Akalin, M.T. 1978. Equipment life and maintenance cost survey. ASHRAE Transactions 84(2):94-106. DOE. International performances measurement and verification protocol. Publication No. DOE/EE-0157. U.S. Department of Energy. Dohrmann, D.R. and T. Alereza. 1986. Analysis of survey data on HVAC maintenance costs. ASHRAE Transactions 92(2A):550-65. Easton Consultants. 1986. Survey of residential heat pump service life and maintenance issues. Available from American Gas Association, Arling- ton, VA (Catalog No. S-77126). Grant, E., W. Ireson, and R. Leavenworth. 1982. Principles of engineering economy. John Wiley and Sons, New York. Haberl, J. 1993. Economic calculations for ASHRAE Handbook. Energy Systems Laboratory Report No. ESL-TR-93/04-07. Texas A&M Univer- sity, College Station, TX. Kreider, J. and F. Kreith. 1982. Solar heating and cooling. Hemisphere Publishing, Washington, D.C. Kreith, F. and J. Kreider. 1978. Principles of solar engineering. Hemisphere Publishing, Washington, D.C. Lippiatt, B.L. 1994. Energy prices and discount factors for life-cycle cost analysis 1993. Annual Supplement to NIST Handbook 135 and NBS Special Publication 709. NISTIR 85-3273.7. National Institute of Standards and Technology, Gaithersburg, MD. Lovvorn, N.C. and C.C. Hiller. 1985. A study of heat pump service life. ASHRAE Transactions 91(2B):573-88. NIST. Annual Supplement to NIST Handbook 135. National Institute of Standards and Technology, Gaithersburg, MD. NIST and DOE. Building life-cycle cost (BLCC) computer program. Avail- able from National Institute of Standards and Technology, Office of Applied Economics, Gaithersburg, MD. OMB. 1972. Guidelines and discount rates for benefit-cost analysis of federal programs. Circular A-94. Office of Management and Budget, Wash- ington, D.C. Riggs, J.L. 1977. Engineering economics. McGraw-Hill, New York. Ruegg, R.T. Life-cycle costing manual for the Federal Energy Management Program. NIST Handbook 135. National Institute of Standards and Tech- nology, Gaithersburg, MD. U.S. Department of Commerce, Bureau of Economic Analysis. Survey of current business. U.S. Government Printing Office, Washington, D.C. USA-CERL and USACE. Life cycle cost in design (LCCID) computer program. Available from Building Systems Laboratory, University of Illi- nois, Urbana. USACE. PC-Econpack computer program. U.S. Army Corps of Engineers, Huntsville, AL. BIBLIOGRAPHY ASTM. 1992. Standard terminology of building economics. Standard E833 Rev A-92. American Society for Testing and Materials, West Consho- hoken, PA. Kurtz, M. 1984. Handbook of engineering economics: A guide for engineers, technicians, scientists, and managers. McGraw-Hill, New York. Quirin, D.G. 1967. The capital expenditure decision. Richard D. Win, Inc., Homewood, IL. CHAPTER 36 TESTING, ADJUSTING, AND BALANCING Terminology 36.1 General Criteria 36.1 Air Volumetric Measurement Methods 36.2 Balancing Procedures for Air Distribution 36.3 Variable Volume Systems 36.4 Principles and Procedures for Balancing Hydronic Systems 36.6 Water-Side Balancing 36.8 Hydronic Balancing Methods 36.9 Fluid Flow Measurement 36.11 Steam Distribution 36.14 Cooling Towers 36.15 Temperature Control Verification 36.15 Field Survey for Energy Audit 36.16 Testing for Sound and Vibration 36.18 Testing for Sound 36.18 Testing for Vibration 36.20 HE system that controls the environment in a building is a Tdynamic entity that changes with time and use, and it must be rebalanced accordingly. The designer must consider initial and sup- plementary testing and balancing requirements for commissioning. Complete and accurate operating and maintenance instructions that include intent of design and how to test, adjust, and balance the building systems are essential. Building operating personnel must be well trained, or qualified operating service organizations must be employed to ensure optimum comfort, proper process operations, and economy of operation. This chapter does not suggest which groups or individuals should perform the functions of a complete testing, adjusting, and balancing procedure. However, the procedure must produce repeatable results that meet the intent of the designer and the requirements of the owner. Overall, one source must be responsible for testing, adjusting, and balancing all systems. As part of this responsibility, the testing organization should check all equipment under field conditions to ensure compliance. Testing and balancing should be repeated as the systems are ren- ovated and changed. The testing of boilers and other pressure ves- sels for compliance with safety codes is not the primary function of the testing and balancing firm; rather it is to verify and adjust operating conditions in relation to design conditions for flow, temperature, pressure drop, noise, and vibration. ASHRAE Standard 111 outlines detailed procedures not covered in this chapter. TERMINOLOGY Testing, adjusting, and balancing is the process of checking and adjusting all the environmental systems in a building to produce the design objectives. This process includes (1) balancing air and water distribution systems, (2) adjusting the total system to provide design quantities, (3) electrical measurement, (4) establishing quantitative performance of all equipment, (5) verifying automatic controls, and (6) sound and vibration measurement. These procedures are accomplished by checking installations for conformity to design, measuring and establishing the fluid quantities of the system as required to meet design specifications, and recording and reporting the results. The following definitions are used in this chapter. Refer to ASH- RAE Terminology of Heating, Ventilation, Air Conditioning, and Refrigeration (1991) for additional definitions. Test. Determine quantitative performance of equipment. Balance. Proportion flows within the distribution system (sub- mains, branches, and terminals) according to specified design quantities. Adjust. Regulate the specified fluid flow rate and air patterns at the terminal equipment (e.g., reduce fan speed, adjust a damper). Procedure. An approach to and execution of a sequence of work operations to yield repeatable results. Report forms. Test data sheets arranged in logical order for sub- mission and review. The data sheets should also form the permanent record to be used as the basis for any future testing, adjusting, and balancing. Terminal. A point where the controlled medium (fluid or energy) enters or leaves the distribution system. In air systems, these may be variable air or constant volume boxes, registers, grilles, diffusers, louvers, and hoods. In water systems, these may be heat transfer coils, fan coil units, convectors, or finned-tube radia- tion or radiant panels. GENERAL CRITERIA Effective and efficient testing, adjusting, and balancing require a systematic, thoroughly planned procedure implemented by experi- enced and qualified staff. All activities, including organization, calibration of instruments, and execution of the actual work, should be scheduled. Air-side must be coordinated with water-side work. Pre- paratory work includes planning and scheduling all procedures, col- lecting necessary data (including all change orders), reviewing data, studying the system to be worked on, preparing forms, and making preliminary field inspections. Leakage can significantly reduce performance; therefore ducts must be designed, constructed, and installed to minimize and control air leakage. During construction, all duct systems should be sealed and tested for air leakage; and water, steam, and pneumatic piping should be tested for leakage. Design Considerations Testing, adjusting, and balancing begin as design functions, with most of the devices required for adjustments being integral parts of the design and installation. To ensure that proper balance can be achieved, the engineer should show and specify a sufficient number of dampers, valves, flow measuring locations, and flow balancing devices; these must be properly located in required straight lengths of pipe or duct for accurate measurement. The testing procedure depends on system characteristics and layout. The interaction between individual terminals varies with pressures, flow requirements, and control devices. The design engineer should specify balancing tolerances. Sug- gested tolerances are ±10% for individual terminals and branches in noncritical applications and ±5% for main ducts. For critical applications where differential pressures must be maintained, the following tolerances are suggested: Positive zones Supply air 0 to +10% Exhaust and return air 0 to −10% Negative zones Supply air 0 to −10% Exhaust and return air 0 to +10% The preparation of this chapter is assigned to TC 9.7, Testing and Balancing. Testing, Adjusting, and Balancing 36.5 Varying Fan Speed Electrically. This method of control, which varies the voltage or frequency to the fan motor, is usually the most efficient. Some versions of motor drives may cause electrical noise and affect other devices. In controlling VAV fan systems, the location of the static pressure sensors is critical and should be field verified to give the most representative point of operation. After the terminal boxes have been proportioned, the static pressure control can be verified by observing static pressure changes at the fan discharge and the static pressure sensor as the load is simulated from maximum airflow to minimum airflow (i.e., set all terminal boxes to balanced airflow conditions and determine whether any changes in static pressure occur by plac- ing one terminal box at a time to minimum airflow, until all terminals are placed at the minimal airflow setting). Care should be taken to verify that the maximum to minimum air volume changes are within the fan curve performance (speed or total pressure). Diversity Diversity may be used on a VAV system, assuming that the total airflow is lower by design and that all terminal boxes will never fully open at the same time. Care should be taken to avoid duct leakage. All ductwork upstream of the terminal box should be considered as medium-pressure ductwork, whether in a low- or medium- pressure system. A procedure to test the total air on the system should be established by setting terminal boxes to the zero or minimum position nearest the fan. During peak load conditions, care should be taken to verify that an adequate pressure is available upstream of all terminal boxes to achieve design airflow to the spaces. Outside Air Requirements Maintaining the space under a slight positive or neutral pressure to atmosphere is difficult with all variable volume systems. In most systems, the exhaust requirement for the space is constant; hence, the outside air used to equal the exhaust air and meet the minimum outside air requirements for the building codes must also remain constant. Due to the location of the outside air intake and the changes in pressure, this does not usually happen. The outside air should enter the fan at a point of constant pressure (i.e., supply fan volume can be controlled by proportional static pressure control, which can control the volume of the return air fan). Makeup air fans can also be used for outside air control. Return Air Fans If return air fans are required in series with a supply fan, the type of control and sizing of the fans is most important. Serious over- and underpressurization can occur, especially during the economizer cycle. Types of VAV Systems Single-Duct VAV. This system incorporates a pressure-dependent or -independent terminal and usually has reheat at some pre- determined minimal setting on the terminal unit or separate heating system. Bypass. This system incorporates a pressure-dependent damper, which, on demand for heating, closes the damper to the space and opens to the return air plenum. Bypass sometimes incorporates a constant bypass airflow or a reduced amount of airflow bypassed to the return plenum in relation to the amount supplied to the space. No economical value can be obtained by varying the fan speed with this system. A control problem can exist if any return air sensing is done to control a warm-up or cool-down cycle. VAV Using Single-Duct VAV and Fan-Powered, Pressure- Dependent Terminals. This system has a primary source of air from the fan to the terminal and a secondary powered fan source that pulls air from the return air plenum before the additional heat source. This system places additional maintenance of terminal fil- ters, motors, and capacitors on the building owner. In certain fan- powered boxes, backdraft dampers are a source of duct leakage when the system calls for the damper to be fully closed. Typical applications include geographic areas where the ratio of heating hours to cooling hours is low. Double-Duct VAV. This type of terminal incorporates two single-duct variable terminals. It is controlled by velocity controllers that operate in sequence so that both hot and cold ducts can be opened or closed. Some controls have a downstream flow sensor in the terminal unit to maintain either the heating or the cooling. The other flow sensor is in the inlet controlled by the thermostat. As this inlet damper closes, the downstream controller opens the other damper to maintain the set airflow. Often, low pressure in the decks controlled by the thermostat causes unwanted mixing of air, which results in excess energy use or discomfort in the space. On most direct digital controls (DDC) inlet control on both ducts is favored in lieu of the downstream controller. Balancing the VAV System The general procedure for balancing a VAV system is 1. Determine the required maximum air volume to be delivered by the supply and return air fans. Diversity of load usually means that the volume will be somewhat less than the outlet total. 2. Obtain fan curves on these units, and request information on surge characteristics from the fan manufacturer. 3. If an inlet vortex damper control is to be used, obtain the fan manufacturer’s data pertaining to the deaeration of the fan when used with the damper. If speed control is used, find the maximum and minimum speed that can be used on the project. 4. Obtain from the manufacturer the minimum and maximum operating pressures for terminal or variable volume boxes to be used on the project. 5. Construct a theoretical system curve, including an approximate surge area. The system curve starts at the minimum inlet static pressure of the boxes, plus system loss at minimum flow, and terminates at the design maximum flow. The operating range using an inlet vane damper is between the surge line intersec- tion with the system curve and the maximum design flow. When variable speed control is used, the operating range is between (a) the minimum speed that can produce the necessary minimum box static pressure at minimum flow still in the fan’s stable range and (b) the maximum speed necessary to obtain maximum design flow. 6. Position the terminal boxes to the proportion of maximum fan air volume to total installed terminal maximum volume. 7. Set the fan to operate at approximate design speed (increase about 5% for a full open inlet vane damper). 8. Check a representative number of terminal boxes. If a wide variation in static pressure is encountered, or if the airflow at a number of boxes is below minimum at maximum flow, check every box. 9. Run a total air traverse with a pitot tube. 10. Increase the speed if static pressure and/or volume are low. If the volume is correct, but the static is high, reduce the speed. If the static is high or correct, but the volume is low, check for system effect at the fan. If there is no system effect, go over all terminals and adjust them to the proper volume. 11. Run steps (7) through (10) with the return or exhaust fan set at design flow as measured by a pitot-tube traverse and with the system set on minimum outdoor air. 12. Proportion the outlets, and verify the design volume with the VAV box on the maximum flow setting. Verify the minimum flow setting. Testing, Adjusting, and Balancing 36.7 Heat Transfer at Reduced Flow Rate The typical heating-only hydronic terminal gradually reduces its heat output as flow is reduced (Figure 1). Decreasing water flow to 50% of design reduces the heat transfer to 90% of that at full design flow. The control valve must reduce the water flow to 10% to reduce the heat output to 50%. The reason for the relative insensitivity to changing flow rates is that the governing coefficient for heat transfer is the air-side coefficient. A change in internal or water-side coefficient with flow rate does not materially affect the overall heat transfer coefficient. This means that (1) heat transfer for water-to- air terminals is established by the mean air-to-water temperature difference, (2) the heat transfer is measurably changed, and (3) a change in the mean water temperature requires a greater change in the water flow rate. A secondary concern also applies to heating terminals. Unlike chilled water, hot water can be supplied at a wide range of temperatures. So, in some cases, an inadequate terminal heating capacity caused by insufficient flow can be overcome by raising the supply water temperature. Design below the temperature limit of 120°C (ASME low-pressure boiler code) must be considered. The previous comments apply to heating terminals selected for a 10 K temperature drop (∆t) and with a supply water temperature of about 93°C. Figure 2 shows the flow variation when 90% terminal capacity is acceptable. Note that heating tolerance decreases with temperature and flow rates and that chilled water terminals are much less tolerant of flow variation than hot water terminals. Dual-temperature heating/cooling hydronic systems are sometimes completed and started during the heating season. Adequate heating ability in the terminals may suggest that the system is balanced. Figure 2 shows that 40% of design flow through the terminal provides 90% of design heating with 60°C supply water and a 5 K temperature drop. Increased supply water temperature establishes the same heat transfer at terminal flow rates of less than 40% design. In some cases, dual-temperature water systems may experi- ence a decreased flow during the cooling season because of the chiller pressure drop; this could cause a flow reduction of 25%. For example, during the cooling season, a terminal that originally heated satisfactorily would only receive 30% of the design flow rate. While the example of reduced flow rate at ∆t = 10 K only affects the heat transfer by 10%, this reduced heat transfer rate may have the following negative effects: 1. The object of the system is to deliver (or remove) heat where required. When the flow is reduced from the design rate, the system must supply heating or cooling for a longer period to maintain room temperature. 2. As the load reaches design conditions, the reduced flow rate is unable to maintain room design conditions. Terminals with lower water temperature drops have a greater tolerance for unbalanced conditions. However, larger water flows are necessary, requiring larger pipes, pumps, and pumping cost. Also, automatic valve control is more difficult. System balance becomes more important in terminals with a large temperature difference. Less water flow is required, which reduces the size of pipes, valves, and pumps, as well as pumping costs. A more linear emission curve gives better system control. Heat Transfer at Excessive Flow The flow rate should not be increased above design in an effort to increase heat transfer. Figure 3 shows that increasing the flow to 200% of design only increases heat transfer by 6% while increasing the resistance or pressure drop 4 times and the power by the cube of the original power (pump laws). Generalized Chilled Water Terminal— Heat Transfer Versus Flow The heat transfer for a typical chilled water coil in an air duct versus water flow rate is shown in Figure 4. The curves shown are based on ARI rating points: 7.2°C inlet water at a 5.6 K rise with entering air at 26.7°C dry bulb and 19.4°C wet bulb. The basic curve applies to catalog ratings for lower dry-bulb temperatures providing a consistent entering air moisture content (e.g., 23.9°C dry bulb, 18.3°C wet bulb). Changes in inlet water temperature, temperature rise, air velocity, and dry- and wet-bulb temperatures will cause terminal performance to deviate from the curves. Figure 4 is only a general representation of the total heat transfer change versus flow for a hydronic cooling coil and does not apply to all chilled water terminals. Comparing Figure 4 with Figure 1 indicates the similarity of the nonlinear heat transfer and flow for both the heating and the cooling terminal. Table 1 shows that if the coil is selected for the load, and the flow is reduced to 90% of the load, three flow variations can satisfy the reduced load at various sensible and latent combinations. Fig. 1 Effects of Flow Variation on Heat Transfer from a Hydronic Terminal (Design ∆t = 10 K and supply temperature = 93°C) Fig. 2 Percent of Design Flow Versus Design for Various Supply Water Temperatures 36.10 1999 ASHRAE Applications Handbook (SI) the desired curve can be determined from the manufacturer’s ratings since these are published as (t ew − t ea ). A second point is established by observing that the heat transfer from air to water is zero when (t ew − t ea ) is zero (consequently, ∆t w = 0). With these two points, an approximate performance curve can be drawn (see Figure 6). Then, for any other (t ew − t ea ), this curve is used to determine the appropriate ∆t w . Example 1. From the following manufacturer certified data, determine the required ∆t w : Capacity = 3 kW t ew = 95°C t ea = 15°C Water flow = 0.1 L/s c p = 4.18 kJ/(kg·K) ρ = 1.0 kg/L Solution: 1. Calculate rated ∆t w . 2. Construct a performance curve as illustrated in Figure 6. 3. From test data: 4. From Figure 6 read ∆t w = 5.4 K, which is required to balance water flow at 0.1 L/s. The water temperature difference may also be calculated as proportion of the rate value as follows: This procedure is useful for balancing terminal devices such as finned tube convectors, where flow measuring devices do not exist and where airflow measurements cannot be made. It may also be used for cooling coils for sensible transfer (dry coil). Flow Balancing by Total Heat Transfer. This procedure deter- mines water flow by running an energy balance around the coil. From field measurements of airflow, wet- and dry-bulb temperatures both upstream and downstream of the coil, and the difference ∆t w between the entering and leaving water temperatures, water flow can be determined by the following equations: (5) (6) (7) where Q w = water flow rate, L/s q = load, W q cooling = cooling load, W q heating = heating load, W Q a = airflow rate, L/s h = enthalpy, kJ/kg t = temperature, °C Example 2. Find the water flow for a cooling system having the following characteristics: Solution: From Equations (5) and (6), The desired water flow is achieved by successive manual adjustments and recalculations. Note that these temperatures can be greatly influenced by the heat of compression, stratification, bypassing, and duct leakage. General Balance Procedures All the variations of balancing hydronic systems cannot be listed; however, the general method should balance the system while minimizing operating cost. Excess pump pressure (excess operating power) can be eliminated by trimming the pump impeller. Allowing excess pressure to be absorbed by throttle valves adds a lifelong operating cost penalty to the operation. The following is a general procedure based on setting the balance valves on the site: 1. Develop a flow diagram if one is not included in the design drawings. Illustrate all balance instrumentation, and include any additional instrument requirements. 2. Compare pumps, primary heat exchangers, and specified terminal units; and determine whether a design diversity factor can be achieved. 3. Examine the control diagram and determine the control adjustments needed to obtain design flow conditions. ∆ t w 3 4.18 0.1 × 1 × 7.18 K== Fig. 6 Coil Performance Curve t ew 80 ° C= t ea 20 ° C= t ew t ea –60 ° C= t ew t ea – () test t ew t ea – () rated ∆ t w () rated ∆ t w () required = 80 20– 95 15– 7.18 × 5.4 K= Test data t ewb = entering wet-bulb temperature = 20.3°C t lwb = leaving wet-bulb temperature = 11.9°C Q a = airflow rate = 10 000 L/s t lw = leaving water temperature = 15.0°C t ew = entering water temperature = 8.6°C From psychrometric chart h 1 = 76.52 kJ/kg h 2 = 52.01 kJ/kg Q w Q 4180⁄∆t w = q cooling 1.20 Q a h 1 h 2 –()= q heating 1.23 Q a t 1 t 2 –()= Q w 1.20 10000 76.52 52.01– ()× 4180 15.0 8.6– () 11.0 L/s== 36.12 1999 ASHRAE Applications Handbook (SI) For example, a manufacturer may test a boiler control valve with 40°C water. Differential pressures from another test made in the field at 120°C may be correlated with the manufacturer’s data by using Equation (8) to account for the density differences of the two tests. When differential heads are used to estimate flow, a density correction must be made because of the shape of the pump curve. For example, in Figure 8 the uncorrected differential reading for pumped water with a density of 900 kg/m 3 is 25 m; the gage conversion was assumed to be for water with a density of 999 kg/m 3 . The uncorrected or false reading gives a 40% error in flow estimation. Differential Head Readout with Manometers Manometers are used for differential pressure readout, especially when very low differentials, great precision, or both, are required. But manometers must be handled with care; they should not be used for field testing because fluid could blow out into the water and rap- idly deteriorate the components. A proposed manometer arrangement is shown in Figure 9. Figure 9 and the following instructions provide accurate manometer readings with minimum risk of blowout. 1. Make sure that both legs of the manometer are filled with water. 2. Open the purge bypass valve. 3. Open valved connections to high and low pressure. 4. Open the bypass vent valve slowly and purge air here. 5. Open manometer block vents and purge air at each point. 6. Close the needle valves. The columns should zero in if the manometer is free of air. If not, vent again. 7. Open the needle valves and begin throttling the purge bypass valve slowly, watching the fluid columns. If the manometer has an adequate available fluid column, the valve can be closed and the differential reading taken. However, if the fluid column reaches the top of the manometer before the valve is completely closed, insufficient manometer height is indicated and further throttling will blow fluid into the blowout collector. A longer manometer or the single gage readout method should then be used. An error is often introduced when converting millimetres of gage fluid to the pressure difference (in kilopascals) of the test fluid. The conversion factor changes with test fluid temperature, density, or both. Conversion factors shown in Table 2 are to a water base, and the counterbalancing water height H (Figure 9) is at room temperature. Orifice Plates, Venturi, and Flow Indicators Manufacturers provide flow information for several devices used in hydronic system balance. In general, the devices can be classified as (1) orifice flowmeters, (2) venturi flowmeters, (3) velocity impact meters, (4) pitot-tube flowmeters, (5) bypass spring impact flowmeters, (6) calibrated balance valves, (7) turbine flowmeters, and (8) ultrasonic flowmeters. The orifice flowmeter is widely used and is extremely accurate. The meter is calibrated and shows differential pressure versus flow. Accuracy generally increases as the pressure differential across the meter increases. The differential pressure readout instrument may be a manometer, differential gage, or single gage (Figure 7). The venturi flowmeter has lower pressure loss than the orifice plate meter because a carefully formed flow path increases velocity head recovery. The venturi flowmeter is placed in a main flow line where it can be read continuously. Velocity impact meters have precise construction and calibration. The meters are generally made of specially contoured glass or plastic, which permits observation of a flow float. As flow increases, the flow float rises in the calibrated tube to indicate flow rate. Velocity impact meters generally have high accuracy. A special version of the velocity impact meter is applied to hydronic systems. This version operates on the velocity pressure difference between the pipe side wall and the pipe center, which causes fluid to flow through a small flowmeter. Accuracy depends on the location of the impact tube and on a velocity profile that cor- responds to theory and the laboratory test calibration base. Gener- ally, the accuracy of this bypass flow impact or differential velocity pressure flowmeter is less than a flow-through meter, which can operate without creating a pressure loss in the hydronic system. The pitot-tube flowmeter is also used for pipe flow measurement. Manometers are generally used to measure velocity pressure differences because these differences are low. The bypass spring impact flowmeter uses a defined piping pressure drop to cause a correlated bypass side branch flow. The side branch flow pushes against a spring that increases in length with increased side branch flow. Each individual flowmeter is calibrated to relate extended spring length position to main flow. The bypass spring impact flowmeter has, as its principal merit, a direct readout. However, dirt on the spring reduces accuracy. The bypass Table 2 Differential Pressure Conversion to Head Fluid Density, kg/m 3 Approximate Corresponding Water Temperature, °C Metre Fluid Head Equal to 1 kPa a 1500 0.680 1400 0.0728 1300 0.0784 1200 0.0850 1100 0.0927 1000 10 0.1020 980 65 0.104 960 95 0.106 940 125 0.108 920 150 0.111 900 170 0.113 800 0.127 700 0.146 600 0.170 500 0.204 a Differential kPa readout is multiplied by this number to obtain metres fluid head when gage is calibrated in kPa. Fig. 9 Fluid Manometer Arrangement for Accurate Reading and Blowout Protection Testing, Adjusting, and Balancing 36.13 is opened only when a reading is made. Flow readings can be taken at any time. The calibrated balance valve is an adjustable orifice flowmeter. Balance valves can be calibrated so that a flow/pressure drop relationship can be obtained for each incremental setting of the valve. A ball, rotating plug, or butterfly valve may have its setting expressed in percent open or degree open; a globe valve, in percent open or number of turns. The calibrated balance valve must be manufac- tured with precision and care to ensure that each valve of a particular size has the same calibration characteristics. The turbine flowmeter is a mechanical device. The velocity of the liquid spins a wheel in the meter, which generates a 4 to 20 mA output that may be calibrated in units of flow. The meter must be well maintained, as wear or water impurities on the bearing may slow the wheel, and debris may clog or break the wheel. The ultrasonic flowmeter senses sound signals, which are calibrated in units of flow. The ultrasonic metering station may be installed as part of the piping, or it may be a strap-on meter. In either case, the meter has no moving parts to maintain, nor does it intrude into the pipe and cause a pressure drop. Two distinct types of ultrasonic meter are available: (1) the transit time meter for HVAC or clear water systems and (2) the Doppler meter for systems handling sewage or large amounts of particulate matter. If any of the above meters are to be useful, the minimum distance of straight pipe upstream and downstream, as recommended by the meter manufacturer and flow measurement handbooks, must be adhered to. Figure 10 presents minimum installation suggestions. Using a Pump as an Indicator Although the pump is not a meter, it can be used as an indicator of flow together with the other system components. Differential pressure readings across a pump can be correlated with the pump curve to establish the pump flow rate. Accuracy depends on (1) accuracy of readout, (2) pump curve shape, (3) actual conformance of the pump to its published curve, (4) pump operation without cavitation, (5) air-free operation, and (6) velocity pressure correction. When a differential pressure reading must be taken, a single gage with manifold provides the greatest accuracy (Figure 11). The pump suction to discharge differential can be used to establish pump differential pressure and, consequently, pump flow rate. The single gage and manifold may also be used to check for strainer clogging by measuring the pressure differential across the strainer. If the pump curve is based on fluid head, pressure differential, as obtained from the gage reading, needs to be converted to head, which is pressure divided by the fluid density and gravity. The pump differential head is then used to determine pump flow rate (Figure 12). As long as the differential head used to enter the pump curve is expressed as head of the fluid being pumped, the pump curve shown by the manufacturer should be used as described. The pump curve may state that it was defined by test with 30°C water. This is unim- portant, since the same curve applies from 15 to 120°C water, or to any fluid within a broad viscosity range. Generally, pump-derived flow information, as established by the performance curve, is questionable unless the following precautions are observed: 1. The installed pump should be factory calibrated by a test to establish the actual flow-pressure relationship for that particular pump. Production pumps can vary from the cataloged curve because of minor changes in impeller diameter, interior casting tolerances, and machine fits. 2. When a calibration curve is not available for a centrifugal pump being tested, the discharge valve can be closed briefly to establish the no-flow shutoff pressure, which can be compared to the published curve. If the shutoff pressure differs from that published, draw a new curve parallel to the published curve. While not exact, the new curve will usually fit the actual pumping circumstance more accurately. Clearance between the impeller and casing minimize the danger of damage to the pump during a no-flow test, but manufacturer verification is necessary. 3. Differential pressure should be determined as accurately as pos- sible, especially for pumps with flat flow curves. 4. The pump should be operating air-free and without cavitation. A cavitating pump will not operate to its curve, and differential readings will provide false results. 5. Ensure that the pump is operating above the minimum net positive suction pressure. 6. Power readings can be used (1) as a check for the operating point when the pump curve is flat or (2) as a reference check when there is suspicion that the pump is cavitating or providing false readings because of air. Fig. 10 Minimum Installation Dimensions for Flowmeter Fig. 11 Single Gage for Differential Readout Across Pump and Strainer Fig. 12 Differential Pressure Used to Determine Pump Flow [...]... functions, while applications software is written to accomplish a particular task that can require many function and utility components HVAC application programs calculate such items as loads, energy, and piping design General-purpose applications software, such as accounting and word processing programs, is discussed in the section on General Productivity Tools Another specialized area of applications. .. and particle transport in cleanrooms The Journal of Environmental Sciences, 31 Rabl, A 1988 Parameter estimation in buildings Methods for dynamic analysis of measured energy use ASME Journal of Solar Energy, Engineering, 110 Sharimugavelu, I., T.H Kuehn, and B.Y.H Liu 1987 Numerical simulation of flow fields in clean rooms Proceedings of the Institute of Enivironmental Sciences, 298-303 1999 ASHRAE Applications. .. failure CHAPTER 38 COMPUTER APPLICATIONS Application Concepts General Productivity Tools Engineering Design Calculations Simulation Programs Graphics Applications 38.1 38.3 38.4 38.8 38.9 T HE use of computers in the heating, refrigerating, and airconditioning industry has come about because of the variety of engineering analysis programs for the HVAC industry, an even... terminals program can accommodate Computer Applications • • • • • Maximum number of circuits program can handle Maximum number of nodes each circuit can have Maximum number of nodes program can handle Compressibility effects for gases and steam Provision for two-phase fluids Acoustic Calculations Chapter 46 summarizes sound generation and attenuation in HVAC applications Applying this data and methodology... manufacturer in the presentation of acoustical data Predictive acoustic software allows designers to look at HVACgenerated sound in a realistic, affordable time frame HVAC oriented acoustic consultants generally assist designers by providing cost-effective sound control ideas for sound-critical applications Refrigerant Properties REFPROP (NIST 1996) is a program that allows the user to examine thermodynamic... including maintenance, software, and personnel costs Software must be The preparation of this chapter is assigned to TC 1.5, Computer Applications Monitoring and Control Applications of Artificial Intelligence Internet ASHRAE Developed Software 38 .10 38.11 38.12 38.13 either bought or developed, and although it may seem expensive to buy, it is almost always more expensive to... Environmental Balancing Bureau, Vienna, VA NEBB 1986 Testing, adjusting, balancing manual for technicians, 1st ed SMACNA 1993 HVAC systems—Testing, adjusting and balancing, 2nd ed Sheet Metal and Air Conditioning Contractors’ National Association, Merrifield, VA SMACNA 1995 HVAC air duct leakage test manual, 1st ed Trane Company 1988 Trane air conditioning manual The Trane Company, LaCrosse, WI CHAPTER... contractors, manufacturers/suppliers, and owners Operation and maintenance of all HVAC& R systems should be considered during the original design of a building Any successful operation and maintenance program must include proper documentation of the design intent and criteria ASHRAE Guideline 4 provides a methodology to properly document HVAC systems Newly installed systems must be commissioned to ensure that... documents the objectives, establishes evaluation criteria, and commits the maintenance department to basic areas of performance, such as prompt response to mechanical failure and attention to planned functions that protect capital investment and minimize downtime or failure response Failure response classifies maintenance department resources expended or reserved to handle interruptions in the operation or... measurements generally have to be made (1) when the specification requires that the sound levels from HVAC equipment only, as opposed to the sound level in a space, not exceed a certain specified level; (2) when the sound level in the space exceeds a desirable level, in which case the noise contributed by the HVAC system must be determined; and (3) in residential locations where little significant background . Equal to 1 kPa a 1500 0.680 1400 0.0728 1300 0.0784 1200 0.0850 1100 0.0927 100 0 10 0 .102 0 980 65 0 .104 960 95 0 .106 940 125 0 .108 920 150 0.111 900 170 0.113 800 0.127 700 0.146 600 0.170 500. 000 17 500 32 500 0.564 18 300 14 100 10 900 7 0 50 000 15 000 35 000 17 500 32 500 0.513 16 700 12 200 9 100 8 0 50 000 15 000 35 000 17 500 32 500 0.467 15 200 10 600 7 600 Total Cash Flow 27. are suggested: Positive zones Supply air 0 to +10% Exhaust and return air 0 to 10% Negative zones Supply air 0 to 10% Exhaust and return air 0 to +10% The preparation of this chapter is assigned

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