Technical guides for owner/manager of an air conditioning system: volume 8. How manufacturers could help the unfortunate energy auditor

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Lester Clark
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1 Technical guides for owner/manager of an air conditioning system: volume 8 How manufacturers could help the unfortunate energy auditor

Team France (Project coordinator) Armines - Mines de Paris Austria Austrian Energy Agency Belgium Université de Liège Italy Politecnico di Torino Portugal University of Porto Slovenia University of

2 Team France (Project coordinator) Armines - Mines de Paris Austria Austrian Energy Agency Belgium Université de Liège Italy Politecnico di Torino Portugal University of Porto Slovenia University of Ljubljana UK Association of Building Engineers BRE (Building Research Establishment Ltd) Welsh School of Architecture Eurovent-Certification Authors of this volume Cleide Aparecida Silva (Université de Liège) Cristian Cuevas (Université de Liège) Jules Hannay (Université de Liège) Jean Lebrun (Université de Liège) Vladut Teodorese (Université de Liège) The sole responsibility for the content of this publication lies with the authors. It does not represent the opinion of the European Communities. The European Commission is not responsible for any use that may be made of the information contained therein. 2

3 Introduction As already demonstrated in other chapters of the present report, the energy audit of a whole HVAC system is not a funny game. On site measurements are unavoidable and cost a lot of time and money. Obviously, in most installations, the highest energy consumers are the fans. The chillers are then coming in second position. The nominal performances of these components are usually well identified in laboratory and in manufacturers catalogues. But their average performances are more questionable in actual conditions of use. From other part, providing that these components are correctly identified and modelled, they may become valuable measuring instruments. The object of this chapter is to show how the components manufacturers could help the auditor by allowing him to make a better use of the information they already have. Using a fan as air flow meter The principle Airflow rates measurements are difficult in existing distribution networks: long enough straight lines are seldom available or accessible and velocity profiles are usually not uniform enough. A large series of measuring points is required and the final accuracy is often disappointing. A much better solution consists in using the fan as an air flow meter. This can be done in two ways: 1) By taking profit of the well known characteristics of the fan and using easy measurements, as supply and exhaust static pressures, rotation speed and/or electrical power; 2) Even better, by measuring just one reference pressure drop at fan supply. Both procedures are illustrated and validated hereafter. Fan characteristics identified on the basis of manufacturer data A fan is currently modelled with the help of similarity variables: flow, pressure and power factors. These factors can be correlated to each other by polynomial expressions. The main output of a fan model can be the airflow rate, expressed in specific value (in kg/s of dry air), as usually in air conditioning. Other outputs can be the different factors, exhaust air speed, total pressure difference, isentropic power and (isentropic) temperature increase across the fan (these two last outputs can be used as checking information). The fan is supposed to be characterised by the diameter of its impeller (scale variable), its (fictitious) exhaust area and the coefficients of the polynomial correlations. Supply air conditions (temperature, pressure and moisture content), rotation speed and static pressure difference are taken as input variables. 3

4 This gives the information flow diagram of Figure 1. Figure 1: Example of fan polynomial model The first equations of this model are built on the basis of the definitions of two (flow and pressure) similarity factors: Two dynamic pressures are considered: one at the exhaust and the other one at the periphery of the impeller: 4

5 The other non-dimensional variables considered are the isentropic effectiveness and the power factor: The three factors are inter-correlated through polynomial laws such as: This model is from long time well validated and easy to tune Polynomials are fitted to manufacturer s performance data. Total and static pressures have to be carefully distinguished at fan exhaust: manufacturers present fan performance in terms of total pressure rise, whereas the measurements are usually made in terms of static pressures. Fan characteristics are generally provided by the manufacturers as data sheets, showing the operating curves of the fan (relationships among the different variables of the system). A typical example of data sheet is presented in Figure 2. This information might be better used if the manufacturers were giving: The experimental points actually available; The correlation equations actually used to generate the curves. Still today, the fan simulation model built by the manufacturer is only offered to important customers as manufacturers of air handling units. The equations are then embedded inside black box selection and simulation software. 5

An example of selection is made on Figure 3; it corresponds to three different rotation speeds.

6 Figure 2 Fan characteristics as presented by the manufacturer Some points can be selected in the diagram of Figure 2, in order to identify the polynomial laws already presented. An example of selection is made on Figure 3; it corresponds to three different rotation speeds. Figure 3 Reference points selected for parameter identification The phi-psi and lambda-psi regression curves identified on these points are presented in Figure 4 and Figure 5. 6

Figure 4 Identification of the phi-psi characteristic Figure 5 Identification of the lambda-psi characteristic A phi psi characteristic appears to be accurate enough for airflow rate measurements,

In the case of fans with forward-curved blades, the pressure rise is relatively insensitive to airflow rate and a more accurate result can be obtained by using the efficiency characteristic.

7 Figure 4 Identification of the phi-psi characteristic Figure 5 Identification of the lambda-psi characteristic A phi psi characteristic appears to be accurate enough for airflow rate measurements, when the fan has backward-curved blades (as in the present case). The airflow rate can be currently determined with an accuracy of about 5 %. In the case of fans with forward-curved blades, the pressure rise is relatively insensitive to airflow rate and a more accurate result can be obtained by using the efficiency characteristic. The electrical consumption is then a better indicator of the flow rate. However, this second approach requires a correct identification of all electrical losses (electrical motor and frequency driver, if any). Experimental validation The fan The validation is performed on the fan whose characteristics were already identified. The fan considered in this study was originally installed in a box, downstream of a radiator (Figure 6). 7

Figure 6 the tested fan Tightness of the fan box A first experimental arrangement (Figure 7) was made in order to verify the tightness of the fan box.

Injected airflow rate and box-ambient over-pressure were simultaneously measured. This allows identifying a fictitious leakage area.

8 Figure 6 the tested fan Tightness of the fan box A first experimental arrangement (Figure 7) was made in order to verify the tightness of the fan box. The tightness test consisted in injecting air inside the box, thanks to an auxiliary fan, after having closed both (supply and exhaust) openings. Injected airflow rate and box-ambient over-pressure were simultaneously measured. This allows identifying a fictitious leakage area. An example of measuring result is shown in Figure 8. In the case considered, the leakage flow rate is estimated to 16.8 g/s, which corresponds to a leakage (fictitious isentropic nozzle throat) diameter of 27.3 mm. Fan box Small fan Figure 7 Experimental identification of the fan box leakage area 8

9 Fan box φ 30/60 mm. P = 604 Pa t amb = 20,5 C HR amb = 54 % P atm = 981 mbar P = 350 Pa Figure 8 Leakage measurement Experimental characterization of the fan The fan system is characterised at different rotation speeds; it s equipped with a frequency inverter working between 8.5 and 50.5 Hz. The experimental arrangement is shown in Figures 9 to 11. Three tests series have been performed: they correspond to three pressure drop characteristics of the air circuit: 1) With coil and with a diaphragm used to measure the flow rate (highest pressure drop) 2) With coil and with a nozzle used to measure the flow rate (medium pressure drop) 3) Without coil and with a nozzle used to measure the flow rate (lowest pressure drop). P rad P af P op t amb HR amb P atm N inv f inv W & m I W & in Figure 9 Experimental characterization of the fan (principle schema) 9

Figure 10 Experimental arrangement (back view) Figure 11 Experimental arrangement

the electrical power supplied to the fan motor.

10 Figure 10 Experimental arrangement (back view) Figure 11 Experimental arrangement (front view) The inverter is provided with direct measurements of the frequency and of the electrical power supplied to the fan motor. The electrical power supplied to the inverter and the fan rotation speeds are also measured in these tests. This makes possible to identify the inverter loss and the frequency sliding of the electric motor. The tests results are presented in Figure 12 and Figure 13. Figure 12 Inverter loss and efficiency 10

global electromechanical loss and corresponding efficiency of the invertermotor subsystem (Figure 14); 2) The same terms for the whole

11 Figure 13 Motor sliding As it can be seen here, both the inverter loss and the motor sliding can be identified through linear correlations. By combining these measuring results with the characteristics already identified (from manufacturer data), it s also possible to identify: 1) The global electromechanical loss and corresponding efficiency of the invertermotor subsystem (Figure 14); 2) The same terms for the whole inverter-motor-fan system (Figure 15). Such characteristics would be easy to combine with on site measurements. Figure 14 Fan shaft power as function of motor electrical power 11

The agreement is considered as satisfactory (the actual total pressure at fan exhaust is not directly measured here, but re-calculated by reference to a hypothetical exhaust area).

12 Figure 15 Global loss and global efficiency The test results were also used to validate the laws identified from manufacturer data. An example of such checking is presented in Figure 16: simulated and measured total pressure differences are compared for same flow rates and rotation speeds. The agreement is considered as satisfactory (the actual total pressure at fan exhaust is not directly measured here, but re-calculated by reference to a hypothetical exhaust area). Figure 16 Comparison among simulated and measured pressure differences Validation of a much more expedient method A much more expedient method is proposed by some manufacturers (as the present one). The fan selected for this study was equipped with openings and connecting pipes for differential pressure measurements as shown in Figure 17 and Figure

Figure 19 Measuring device (inside the fan) The measuring

13 Figure 17 Differential pressure measuring device (principle) Figure 18 Measuring device (outside the fan) Figure 19 Measuring device (inside the fan) The measuring principle is described in the manufacturer catalogue as follows: 13

This means that the fan inlet is used as a measuring nozzle. The K constant contains the flow contraction effect. The method can be validated with the test results available.

Figure 20 Exhaust diameter of the (fictitious) fan supply nozzle A (fairly constant) fictitious diameter of 193 mm is here identified. The actual diameter is 257 mm.

14 This means that the fan inlet is used as a measuring nozzle. The K constant contains the flow contraction effect. The method can be validated with the test results available. A fictitious nozzle exhaust diameter is identified at each regime. Examples of results are presented in Figure 20. Figure 20 Exhaust diameter of the (fictitious) fan supply nozzle A (fairly constant) fictitious diameter of 193 mm is here identified. The actual diameter is 257 mm. This corresponds to a contraction factor of (193/257) 2 = and is in good agreement with the K value indicated by the manufacturer. Using refrigeration compressor as enthalpy flow meter The principle Determining on site the cooling power actually provided by a chiller is also a delicate matter. That power should correspond to the enthalpy flow of the secondary fluid (usually water or brine) supplying the evaporator. But neither the flow rate, nor the supply-exhaust temperature difference are easy to measure. The measurements are not easier on refrigerant side. 14

15 An interesting alternative consists in using the chiller compressor as enthalpy flow meter. But, as seen hereafter, this would require a better dialog with manufacturers Data provided by the manufacturers Analysis Each manufacturer seems having his own way to present data in his catalogue. Unfortunately, this personal presentation can be a source of confusion. Data available are nor fully clear, neither complete and the reader is, most of the time, recommended to contact the local manufacturer office as solution. Unfortunately, these offices are not always able to help: it may occur that the information is just no more available, because of too old machines, no more produced or replaced by new models. Some examples of machine designations and physical data presentations used by some manufacturers are presented hereafter, outlining the differences found among them. Examples Figure 21 gives the designation data found in the catalogue data of manufacturer A. Manufacturer A Evaporator cooling capacity Figure 21 Chiller designation of manufacturer A. Figure 22 gives the designation data found in the catalogue of manufacturer B. 15

16 Reference evaporator cooling capacity Nominal Tons Figure 22: Chiller designation of manufacturer B The presentation of physical data is not standardized, as seen in Figure 21 and Figure 22. Here also, each manufacturer has his personalised presentation and attention must be paid to each item presented (nominal loads, refrigerant type, fan capacity, flow rates, etc.), when manipulating catalogues coming from different manufacturers. The units must be carefully identified: Some manufacturers use a semi- SI system with kw, bar, kg/s, m³/h, etc. Other ones are still using Imperial units. But other elementary questions must be answered, as, for example, if the data refers to one, or to several devices (e.g. one or two condenser fans), to partial or global air flow rate, etc. 16

17 Attention! Figure 23 Chiller physical data given by manufacturer A 17

Figure 24 Chiller, physical data given by manufacturer B ARI and Eurovent standards Hopefully, standardized ARI and/or Eurovent reference data are usually also available

18 Figure 24 Chiller, physical data given by manufacturer B ARI and Eurovent standards Hopefully, standardized ARI and/or Eurovent reference data are usually also available in main manufacturer catalogues and can be freely downloaded from Internet. This is a great advantage for the users. There exist other standards, but they are not free! 18

19 Chiller modelling and parameter identification Modelling Various simulation models are currently available. An example of information flow diagram used for chiller modelling is presented in Figure 22. It corresponds to the model of a chiller with scroll compressor(s) and with air-cooled condenser. In this model, the condenser and the evaporator are modelled as fictitious semi-isothermal heat exchangers. A hypothetical proportional control law is applied to the condenser fan. Figure 25 Example information flow diagram The same model can be used in two steps: 1) Parameter identification on one or several reference points; 2) Simulation in all other conditions of use. Attention must be paid to the domain of validity of such model, after tuning: according to the manufacturer: no extrapolation should be done outside the domain covered by the catalogue. Parameter identification This is most delicate operation. If well done, simulation is no more a problem. Figure 26 shows the data usually found in the manufacturer catalogue. A block diagram of the identification procedure is presented in Figure 27. The parameters identified are indicated in Figure 28. The identification process can be done manually and iteratively, by considering the result trends after each step. Default values are used as first guesses for each component separately at the nominal point. These values are tuned in order to obtain results that fit to all orders of magnitude. Finally the parameters are tuned again, in order to obtain a better agreement with all experimental results and/or all manufacturer data available. 19

20 The parameter identification is considered as satisfactory if all thermal and electrical powers are predicted with accuracy of the order of ± 2 %. 20

cd condenser ev evaporator cp compressor e v expansion valve taex_cd tex_cd Air Condenser tsu_cd tsu_e v tasu_cd Compressor taex_cp Expansion valve tex_e v Manufacturer data Q & ev

21 cd condenser ev evaporator cp compressor e v expansion valve taex_cd tex_cd Air Condenser tsu_cd tsu_e v tasu_cd Compressor taex_cp Expansion valve tex_e v Manufacturer data Q & ev W & tsu_cp cp tsu_ev Evaporator tex_ev R22 twex_ev twsu_ev M & w Figure 26 Data usually found in a manufacturer catalogue Figure 27 Block diagram of the identification procedure 21

proper data, in other to calculate the chiller

22 Figure 28 Parameters identified for chiller model Simulation At this stage, any user can introduce his proper data, in other to calculate the chiller performances, as shown in Figure 29. Figure 29 Chiller performances calculation 22

23 Examples of results The examples of results presented hereafter are obtained with data provided by both manufacturers (A and B) already selected. The agreement between simulation and catalogue data is very satisfactory Q ev Q ev,ari Q ev,part-load,50% W cp W cp,ari W cp,part-load,50% Q ev [W] Wcp [W] Q ev,man [W] W cp,man [W] Figure 30 Calculated versus catalogue data (manufacturer A) Q ev Q Qev,ARI W W ARI Q ev [W] W [W] Q ev,man [W] W man [W] Figure 31 Calculated versus catalogue data (manufacturer A) Illustration of the measuring method by simulation The compressor model (contained in the chiller model already presented) can be used for an easy and accurate determination the chiller cooling power on site. The two examples of simulation results presented in Figure 32 and Figure 33 demonstrate that very simple (linear) relationships could be used to determine the chiller cooling power as function of the refrigerant pressure measured at compressor supply. In full load and at constant rotation speed, this supply pressure is the almost unique variable to be considered. A slight shift can be applied to this law as function 23

Figure 33 Chiller cooling power as function of the compressor supply pressure (with compressor exhaust pressure as second independent variable) Such

24 of a second variable: the temperature of the secondary fluid at condenser supply (Figure 32) or, more directly, the condensing pressure (Figure 33). Figure 32 Chiller cooling power as function of the compressor supply pressure (with condenser supply air temperature as second independent variable) Figure 33 Chiller cooling power as function of the compressor supply pressure (with compressor exhaust pressure as second independent variable) Such procedure is easy to apply on site, because, most of the time, both (evaporation and condensation) pressures are actually given on the chiller control board 24

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