Refrigeration and Air Conditioning 3 E Part 13 ppsx

354 Refrigeration and Air-Conditioning 1. Incorrect setting of head pressure controls 2. Dirty or choked spray nozzles in water tower or evaporative condenser, so that the surface is not fully wetted 3. Non-condensible gas in circuit 4. Bad location of condensers, so that air recirculates 5. Undersized condensers 6. Dirty fins on air-cooled condensers 7. Fans not working or broken 8. Water strainers blocked 9. Undersize pumps fitted 10. Air in water circuit While all these factors affect the good running of positive-displace- ment compressors, the effect is far worse with centrifugal machines, which can approach stall condition and so give a much reduced cooling duty. 34.7 Maintenance The good running order of equipment depends on the standards of maintenance. This is a running cost to be assessed with all the others. If it is found to be faulty, the investigation must consider what this is costing in terms of plant inefficiency and the expenditure to reach acceptable standards. This might be in the replacement or extra training of staff, or in contracting the work out. If the latter, the cost must include supervisory expenses. 34.8 Remedial action on existing equipment The faults described above are largely self-revealing and most of them can be corrected or improved without a great deal of expen- diture. The presence of separate metering devices should give an immediate indication of the savings made. 34.9 Improved controls and equipment on existing plant Deficiencies on the original plant might be corrected by comparatively minor improvements, changes and additions. Each should be assessed for its individual contribution to energy economy and how it may improve the performance of other parts of the system. 1. Optimum-start controls. 2. Ambient-biased set point controls. Efficiency and economy in operation 355 3. Modifications to give improved air and water flows, where these were shown to be deficient, i.e. increase fan speeds, change fans, change pumps, improve ductwork or piping to reduce pressure losses. 4. Improved defrost control, to defrost coils only when and for as long as necessary. 5. Improved cold store door-operating mechanisms (see Figure 15.8). 6. Improved condenser pressure control. If the expansion valve is too tightly rated to accept lower condenser pressure, change the expansion valve, possibly for the electronic type. 7. Automatically switch off plant which is not in use (boiler in summer, tower in winter, lights at night, etc.). 8. Switch off some of the cold store fans and coolers at night and weekends. 9. Fit an automatic load-shedding maximum demand limiter. 10. Resite condensers for better air flow. More drastic items may be: 1. Replace worn, obsolete or undersize compressors, evaporators or condensers. 2. Add new compressors, evaporators or condensers if these can be shown to be economical. 34.10 Design of systems for energy economy Previous chapters have outlined the methods of estimating loads, choosing methods to achieve the required conditions, and how to select and balance plant for correct operation. They have also men- tioned the factors which will give economy in running costs. The maximum use should be made of energy-saving methods, where these may be applicable. Some of these are: 1. Use of all fresh air for air-conditioning, if required in cold weather 2. Provide mid-season heating from condenser heat or heat pump (reverse-cycle) operation 3. Run plant at night on low-cost electricity and make ice, to use for chilled water when load comes on (ice-bank) 4. Switch plant off for periods when electricity is at a premium tariff 5. Two-speed or electronically speed-controlled motors for lower compressor, fan and pump speeds at low load 6. Arrange the coolers within a cold store so that they will give adequate air circulation at night when half of them are switched off. 356 Refrigeration and Air-Conditioning Much attention has been given in recent years to the power consumed in the refrigeration process and the development of more efficient compressors. A few points to consider are: 1. Avoid high compression ratios on piston compressors. 2. Avoid single-stage compression for very low temperatures. 3. Avoid machines which are working at the upper or lower limits of their range. 4. Always ask the running power required at load conditions. The resulting system design will not be the lowest in first cost. 34.11 Commitment to energy savings A positive energy policy needs to be a company decision, taken at boardroom level and backed by boardroom authority, since it cuts across departmental boundaries and may conflict with the opinions of senior staff. Typical objections are: 1. The capital, operating, maintenance and fuel costs come from four separate budgets, possibly accounted for by four different managers, so these budgets need to be adjusted. Separate fuel meters are needed to prove the savings, which might otherwise be held in question. 2. There may be some disruption to normal working while the schemes are being carried out. This may affect departments not concerned directly with the programme. 3. Staff may need to be released for training schemes. 4. The improvements may need changes in operating techniques which are thought to be adequate already. Some of these, such as the tightening of discipline of fork-lift truck drivers, may provoke open conflict, which must be foreseen and headed off. It is important to be able to quantify the results of the energy programme and make these known to all concerned. A conservation programme of this sort is an ongoing process and should keep all staff concerned alert to the possibilities of further improvements. 35 Catalogue selection 35.1 General Manufacturers will publish rating and application data for their products, based on standard test conditions and for the more usual range of uses. They cannot be expected to have accurate figures for every possible combination of conditions for an individual purpose, although most will produce estimates if asked. The widespread use of packaged units of all sizes requires interpretation of catalogue data by applications engineers, sales engineers, and others, and by the end user. The first step is to be certain of the basis of the published data and consider in what ways this will be affected by different conditions. Revised figures can then usually be determined. For extensive interpretation work, simple mathematical models of performance can be constructed [69]. 35.2 Compressors Refrigeration compressors which will probably be used on flooded evaporators (R.717 and the larger machines generally) will be rated with the suction at saturated conditions, since there will be little or no superheat in practice. Compressors for dry expansion systems may be rated at a stated amount of superheat, commonly 8 K. There will be a pressure drop and heat gain in the suction line, and these are frequently ignored if the pipe run is short. In other cases, some allowance must be made. Both these factors will increase the specific volume. Example 35.1 An ammonia compressor is rated at 312 kW with saturated suction at –15°C. It is installed with a very long suction line, causing a pressure drop of 0.4 bar, and picks up 6 K superheat from its evaporator condition. Estimate the capacity loss. 358 Refrigeration and Air-Conditioning Evaporator pressure at – 15°C = 2.36 bar abs. Suction pressure, 2.36 – 0.4 = 1.96 bar abs. Rating suction temperature = – 15.0°C Actual suction temperature, – 15 + 6 K = – 9.0°C The absolute gas pressures must be used in this calculation (see Section 1.4). The volume pumped by the compressor will remain about the same, but the density of the gas is reduced, and thence the mass flow. Using the General Gas Laws: m m pT pT 2 1 21 12 = = 1.96 258.15 2.36 264.15 = 0.81 × × So the capacity loss is of the order of 19%, or 59 kW. There may also be a slight drop due to the higher compression ratio, ignored here as the condensing pressure is not known. Halocarbon systems are almost invariably controlled by mechanical or electronic thermostatic expansion valves, requiring a superheat signal to operate the control. The superheating of the suction gas into the compressor will cause it to expand, resulting in a lower mass flow for a given swept volume. Reduction of the superheat setting of the expansion valve will therefore result in better use of the compressor. The limit will be reached when there is insufficient signal to work the expansion valve. Example 35.2 An R.22 compressor is rated at 15.9 kW when eva- porating at – 5°C, with 8 K superheat. Estimate the gain in capacity if it can be run safely with half the superheat. Rating suction temperature, – 5 + 8 = 3°C = 276.15 K Working suction temperature, – 5 + 4 = – 1°C = 272.15 K Ratio of mass pumped = m m T T 2 1 1 2 = + 276.15 272.15 = 1.015 This gives a gain in capacity of about 1.5%, or 0.24 kW. There will also be a gain in usage of the evaporator coil and a corresponding rise in the evaporator temperature, giving a further increase in compressor capacity. This would need to be evaluated from the compressor curves, but might be a further 1%. Catalogue selection 359 Example 35.3 A hermetic compressor is rated at 18.2 kW for R.22 when evaporating at 7°C, suction superheated to 35°C, condensing at 54°C, and with 8 K subcooling of the liquid. Assuming that the inlet gas picks up another 30 K as it passes over the compressor motor, estimate the change in capacity if the suction is superheated to 12°C. (a) Change in mass flow: Compressor inlet temperature, rating, 35 + 30 = 65°C = 338.15 K actual, 12 + 30 = 42°C = 315.15 K m m T T 2 1 1 2 = = 338.15 315.15 = 1.073 (b) Change in enthalpy (kJ/kg): Enthalpy of suction gas at 35°C = 329.8 Enthalpy of suction gas at 12°C = 000.0 311.7 Enthalpy of liquid at (54 – 8) 46°C = 157.0 157.0 Refrigerating effect (kJ/kg) = 172.8 154.7 Change in enthalpy, 154.7 172.8 = 0.895 Overall change in capacity, 1.073 × 0.895 = 0.96 Corrected working capacity, 18.2 × 0.96 = 17.5 kW 35.3 Condensing units Rating curves for condensing units (see also Section 13.2) will be for stated entering temperatures of the condensing medium – air or water. These may not go as high as the particular application may demand, and figures must be extrapolated. The main effects of a higher condensing temperature will be a drop in the refrigerating effect, since the liquid enters the expansion valve hotter, and a decrease in volume pumped due to the lower volumetric efficiency. There will also be an increase in the drive motor power. Example 35.4 An air-cooled condensing unit is rated at 13.7 kW on R.22 when evaporating at 5°C and with ambient air at 43°C. Estimate the duty with ambient air at 52°C. 360 Refrigeration and Air-Conditioning Some assumptions must be made regarding the condenser coil performance, and this may have a ∆T of 14 K between the entering air and condensing refrigerant and subcooling the liquid 5 K, with suction gas entering the compressor with 6 K superheat. Rating Working Rating condensing temperature, 43 + 14 = 57 °C Working condensing temperature, 52 + 14 = 66 °C Enthalpy of suction gas at (5 + 6) = 11°C = 312.1 312.1 Enthalpy of liquid at (57 – 5) = 52°C = 165.3 Enthalpy of liquid at (66 – 5) = 61°C = 178.5 Refrigerating effect (kJ/kg) = 146.8 133.6 In addition, the compression ratio has increased considerably and there must be a correction for loss of volumetric efficiency. Rating Working Suction pressure (bar abs) at 5°C = 5.82 5.82 Discharge pressure at 57°C = 22.84 Discharge pressure at 66°C = 27.76 Compression ratio = 3.92 4.77 Volumetric efficiency (from Figure 2.8) = 0.75 0.68 Estimated new duty = 13.7 × 133.6 146.8 0.68 0.75 × = 11.3 kW This is approximate, but probably within 0.2 kW. 35.4 Evaporators The rating of an evaporator will be proportional to the temperature difference between the refrigerant and the cooled medium. Since the latter is changing in temperature as it passes over the cooler surface (see Section 1.8), an accurate calculation for a particular load is tedious and subject error. To simplify the matching of air-cooling evaporators to condensing units, evaporator duties are commonly expressed in basic ratings (see Figure 35.1), in units of kilowatts per kelvin (formerly in British thermal units per hour per degree Fahrenheit). This rating factor is multiplied by the ∆T between the entering air and the refrigerant. Example 35.5 An air-cooling evaporator has a mass air flow of 8.4 kg/s and a published ‘rating’ of 3.8 kW/K. What will be its rated Catalogue selection 361 duty at – 15°C coldroom temperature with refrigerant at –21°C? What is the true ln MTD? Entering air temperature = –15°C Refrigerant temperature = –21°C ‘Rating’ ∆T =6 K Rated duty = 3.8 × 6 = 22.8 kW Reduction in air temperature = 22.8 1.006 8.4× = 2.73 K Air leaving temperature = –15 –2.73 = –17.73°C ln MTD = 6 – 3.27 ln (6/3.27) = 4.5 K It follows that there would be an error at other conditions and the basic rating is only accurate at one point, so this short-cut factor must only be used within the range specified by the manufacturer. The method of balancing such an evaporator with a condensing unit is graphical. The condensing unit capacity is shown as cooling duty against evaporator temperature, line CD in Figure 35.2. The coil rating is plotted as the line AB, with A at the required coldroom (or ‘air-on’) temperature, and the slope of the line AB corresponding to the basic rating. The intersection of this line with the condensing unit curve CD gives the graphical solution of the system balance point. Similar constructions for higher condenser air conditions (EF, GH) or different room temperatures (A 1 B 1 ) will show balance points for these conditions. The graph also indicates the change in evaporating temperature Rating temperature difference 6 K Evaporation –21°C Air in –15°C In MTD 4.5 K Air out –17.73°C Figure 35.1 Basic rating and ln MTD 362 Refrigeration and Air-Conditioning and coil duty when the ambient is lower or higher than the design figure. This will show if there is any necessity to control the evaporating temperature in order to keep the correct plant operation. (See also Sections 9.8 and 9.11.) 35.5 Reduction of air flow Frequently, coil data will be available for a design air flow, but the system resistance reduces this flow to a lower value. There is a double effect: the lowering of the ln MTD and the lower heat transfer from the coil by convection. The outer surface coefficient is the greatest thermal resistivity (compared with conduction through the coil material and the inside coefficient), and a rough estimate of the total sensible heat flow change can be made on the basis of [5] and [6]: h = constant × (V ) 0.8 Example 35.6 An air cooling coil extracts 45 kW sensible heat with air entering at 24°C and leaving at 18°C, with the refrigerant evaporating at 11°C. Estimate the cooling capacity at 95, 90 and 85% mass air flow. Design mass air flow = 45 1.02 (24 – 18) = 7.35 kg/s × An approximate analysis comes out: B B 1 D H F 25°C 35°C 30°C    Air onto condenser C E G AA 1 – 40 –35 –30 – 25 – 20 –15 Evaporating temperature (°C) 30 25 20 15 10 5 0 Cooling capacity (kW) Figure 35.2 Graphical balance of evaporator with condensing unit Catalogue selection 363 Air flow (%) 100 95 90 85 Mass air flow (kg/s) 7.35 6.99 6.62 6.25 Air temperature on coil (°C) 24 24 24 24 ∆ T for 45 kW (K) 6 6.3 6.7 7.1 Air temperature off coil (°C) 18 17.7 17.3 16.9 In MTD, refrigerant at 11°C (K) 9.7 9.5 9.2 9.0 h , in terms of design (from V 0.8 ) (%) 100 96 92 88 Capacity, (45 × h × ln MTD)/9.7 (kW) 45 42.3 39.3 36.7 This first estimate for the evaporator coil performance must now be corrected for the change in compressor duty if it is a direct expansion coil, or of water temperature change if using chilled water. Another method is to re-calculate the basic rating figures at the new air flows and plot these against compressor curves. With all calculations involving convective heat transfer, it must be remembered that the figures are predictions based on previous test data, and not precise. 35.6 Room air-conditioners The catalogue-rated cooling capacity of a room air-conditioner, if not qualified, will be based on ASHRAE Standard 16-1983. This specifies test conditions of air onto the evaporator at 80°F dry bulb, 50% relative humidity (26.7°C, 49.1% saturation), and air onto the condenser at 95°F dry bulb, 75°F wet bulb (35°C and 23.9°C). The original basis for this specification was the ambient condition prevailing in the mass-market area of the USA. For these units, British Standard 2582: Part 1, 1982 gives three sets of alternative rating conditions, corresponding to the ASHRAE Standard, for tropical, arid and temperate ambients. They are: Room air temperature Outside air temperature DB WB DB WB Condition A 27 19 35 24 Condition B 29 19 46 24 Condition C 21 15 27 19 and catalogue ratings quoting BS.2852 will be qualified with the appropriate conditions letter. The International Document ISO R 859 evolved from existing national standards and does not specify any test conditions, only [...]... Enquiries can now go out for equipment to satisfy the need, based on the options presented No attempt should be made to reach a decision until these have been evaluated Appendix Units of measurement The International System of Units (SI) provides a coherent system of measurement units, and all the physical quantities required for refrigeration and air- conditioning can be derived from the basic standards:... Length Mass Time Electric current Temperature Electric potential metre kilogram second ampere kelvin volt m kg s A K V From these basic units are derived: Area Volume Liquid volume Power Force Energy (Work) Pressure also Temperature square metre cubic metre litre watt newton joule pascal bar degree Celsius m2 m3 m3 × 10 3 W (ampere volt) N (kg m/s2) J (N m or W s) Pa (N/m2) bar (Pa × 105) °C (K – 237 .15)... ASHRAE, Atlanta, Georgia, 1988 17 HUNDY, G F., The development of the single screw compressor and oil reduced operation Proceedings of the Institute of Refrigeration, London, April, 1982 18 KVALNES, D E. , The sealed tube test for refrigeration oils ASHRAE Transactions, 1965 37 0 Refrigeration and Air- Conditioning 19 American Society of Heating, Refrigerating and Air- Conditioning Engineers, Systems Handbook,... OUGHTON, R J., Legionnaires’ Disease in refrigeration and associated equipment Proceedings of the Institute of Refrigeration, London, April, 1987 28 Chartered Institution of Building Services Engineers, Minimising the Risk of Legionnaires’ Disease, Technical Memorandum 13, CIBSE, London, 1988 28a ASHRAE Legionellosis Position paper update www.ashrae.org (internet) 28b CIBSE Guide TM 13 update due out in March... kPa Other terms not given here may be encountered from time to time and will be found in standard reference works [1, 2, 4, 10] References 1 American Society of Heating, Refrigerating and Air- Conditioning Engineers, Handbook of Fundamentals, ASHRAE, Atlanta, Georgia, 1985 2 Chartered Institution of Building Services Engineers, Guide Book A, CIBSE, London, 1986 3 DIAMANT , R M E , Insulation Deskbook,... and Air- Conditioning Product cooling load? Heat leakage, sensible and latent? Convection heat gains, sensible and latent? Internal heat gains? Time required? 3 Constraints 4 5 6 7 8 Degree of reliability? Position of plant? Automatic/manned? Refrigerant? Same type of equipment as existing? Possible methods Direct expansion? Indirect – what medium? Part by tower water or ambient air? Thermal storage?... existing installation or the factory Where the standard is for compliance with a safety requirement, a certificate to this effect should be provided, and may be demanded by insurers 35 .8 Analytical catalogue selection Since a large proportion of refrigeration and air- conditioning equipment will be bought on the basis of catalogue data, an analytical approach should be adopted to ensure correct selection... The principles to be applied are those of value analysis – to start with the basic need and no preconceived method, to consider all the different methods of satisfying the need, and to evaluate each of these objectively before moving towards a choice The details of such an approach will vary considerably, and the following guidelines should be taken as an indication of the factors to be considered,... Heat and enthalpy gains through cold room doorways Proceedings of the Institute of Refrigeration, December, 1975 50 MILLER, H W and GORDON BROWN, T P., Recent developments in ground freezing Proceedings of the Institute of Refrigeration, November, 1967 51 Trane Air Conditioning Manual, The Trane Company, LaCrosse, Wi., 1987 52 JONES, W P., Air Conditioning Engineering, Edward Arnold, London, 19 73 53. .. Detectors, 32 4 Dew point, 230 Dewaxing of oils, 57 Document, commissioning, 33 6 Display, refrigerated, 211 Doors, cold store, 182 Dough retarding, 2 03 Dry bulb, 230 Dry expansion, 60 Driers, 116 Dry coolers, 81 Dual duct, 30 3 Ducts, 2 83, 296 Economy of operation, 35 2 Effectiveness, 11 Efficiency, volumetric, 19 Ejector, steam, 26 Emissivity, 11 Energy savings, 35 6 Energy targets, 35 1 Enthalpy, 1 Erection, 131 . faults described above are largely self-revealing and most of them can be corrected or improved without a great deal of expen- diture. The presence of separate metering devices should give an immediate. accounted for by four different managers, so these budgets need to be adjusted. Separate fuel meters are needed to prove the savings, which might otherwise be held in question. 2. There may be some. condensing temperature will be a drop in the refrigerating effect, since the liquid enters the expansion valve hotter, and a decrease in volume pumped due to the lower volumetric efficiency. There will