Study for the D.G. Transportation-Energy (DGTREN) of the Commission of the E.U.

Energy Efficiency and Certification of Central Air Conditioners (EECCAC)

FINAL REPORT - APRIL 2003 VOLUME 3

CO-ORDINATOR: Jérôme ADNOT, ARMINES, France Assisted by Paul WAIDE

PW Consulting, UK PARTICIPANTS Jérôme ADNOT, Philippe RIVIERE, Dominique MARCHIO,

Martin HOLMSTROM, Johan NAESLUND, Julie SABA Centre d’Energétique, Ecole des Mines de Paris, France

Sule BECIRSPAHIC Eurovent Certification Carlos LOPES ADENE-CCE, Portugal Isabel BLANCO IDAE, Spain

Luis PEREZ-LOMBARD, Jose ORTIZ AICIA, Spain

Nantia PAPAKONSTANTINOU, Paris DOUKAS University of Athens, Greece

Cesare M. JOPPOLO Politecnico di Milano, Italy Carmine CASALE AICARR, Italy Georg BENKE EVA, Austria Dominique GIRAUD INESTENE, France

Nicolas HOUDANTEnergie Demain, France

Philippe RIVIERE, Frank COLOMINES Electricité de France Robert GAVRILIUC, Razvan POPESCU, Sorin BURCHIU UTCB, Bucharest Bruno GEORGES ITF, France Roger HITCHIN BRE, UK With the additional participation of experts from Eurovent Cecomaf

© 2003 ARMINES

ARMINES 60, bd St Michel 75272 Paris Cedex 06 France

Tel: (+33) 1 40 51 91 74 Fax: (+33) 1 46 34 24 91 E-mail: jerome.adnot@cenerg.ensmp.fr

All rights reserved, including that of translation into other languages. No part of this publication may be reproduced or transmitted in any form or by any means, electronic or mechanical, including photocopying, recording or any information storage and retrieval system, without permission in writing from ARMINES.

Editorial content: Although great care has been taken in compiling and checking the information given in this publication to ensure that it is accurate, ARMINES shall not be held responsible for the continued currency of the information or for any errors, omissions or inaccuracies in this publication.

CONTENTS

6. TECHNICAL AND ECONOMIC EVALUATION OF THE ELEMENTARY EQUIPMENT USED IN CAC........................................................................... 9

6.1 Energy-engineering analysis of chillers.......................................................................................... 9 Chiller prices as a function of the refrigerating fluid and EER............................................................ 9 Role of condensing medium ................................................................................................................ 9 Additional costs for reversibility........................................................................................................ 10 Defining chiller part-load efficiency.................................................................................................. 10 Available data and simulation tools ................................................................................................... 11 Incremental costs as a function of efficiency ..................................................................................... 11 Optimisation of the chiller used as baseline without any system consideration ................................ 12 Optimisation of a chiller in a system ................................................................................................. 12 Water cooled chillers ......................................................................................................................... 13

6.2 Engineering approach of the performance of Packaged units ................................................... 14 The US energy engineering analysis.................................................................................................. 15 Life cycle cost analysis ...................................................................................................................... 16

6.3 Energy Efficiency of Air Handling Units seen as tradable goods ........................................... 18 Fans integrated in AHU ..................................................................................................................... 18 Heat recovery section of AHU........................................................................................................... 19

7. TECHNICAL & ECONOMIC EVALUATION OF CAC SYSTEM PERFORMANCE AS A FUNCTION OF THE DESIGN OF THE AC SYSTEM .................................................................. 21

7.1 Comparison of different CAC systems......................................................................................... 21 Energy consumption for a given comfort level .................................................................................. 21 Comparison of costs and sensitivities ................................................................................................ 22

7.2 The improvement of the efficiency of air handling systems in CAC ......................................... 22 Primary Air and ventilation ............................................................................................................... 22 Heat recovery on primary air ............................................................................................................. 23 Motors and fans efficiency................................................................................................................. 23 Variable air flow and lower head losses ............................................................................................ 23 Terminal reheat issues........................................................................................................................ 24 Air Side Free Cooling (Economiser) ................................................................................................. 24 Quality of Air Diffusion..................................................................................................................... 25 AHU improvement ............................................................................................................................ 25

7.3 Other cost & efficiency trade-offs................................................................................................. 25 Water-side efficiency by sizing and control....................................................................................... 25 Design of flow in water circulation.................................................................................................... 26 Influence of terminal equipment ........................................................................................................ 27 Simultaneous demand of heating and cooling.................................................................................... 27 Heat rejection..................................................................................................................................... 28

7.4 The possible strength of regulatory efforts and the minimum LCC solutions ......................... 28 Concentration of efforts on Air based systems .................................................................................. 28 The result of optimisation .................................................................................................................. 28

7.5 The possible effects of technical scenarios ................................................................................... 29 Scenario 1 MOVING ALL COOL GENERATORS TO AVERAGE PERFORMANCE................. 30 Scenario 2 THE BEST CHOICE OF COOL GENERATORS FOR THE CUSTOMER BASED ON FULL LOAD INFO ........................................................................................................................... 30 Scenario 3 BAT- THE BEST CONSUMER CHOICE WITH PROPER PART LOAD INFO ......... 31 Scenario 4 GENERALISED FREE COOLING................................................................................. 32

Scenario 5 Generalised Heat recovery ............................................................................................... 32 There is a reduction in heating and cooling demand which is climate dependant. However the head losses increase significantly with our calculation when there is full Air Conditioning equipment in place. As a result of our calculation, DOE2 results are not positive results for Summer. There is a benefit of Heat Recovery in Winter only, whichever the heating system. Heat Recovery is a positive option to be recommended when it is associated to the ventilation system in Winter and can be positively substituted by free cooling in Summer. ............................................................................. 32 Scenario 6 British regulation on AC – heating, cooling and air movement- adapted for each EU climate................................................................................................................................................ 32 Scenario 7 Portuguese regulation on AC – heating, cooling and air movement- adapted for each EU climate................................................................................................................................................ 34 Scenario 8 US regulation on AC – heating, cooling and air movement- adapted for each EU climate........................................................................................................................................................... 34 Scenario 9 Obligation of using Class 2 motors in any part of A/C equipment .................................. 34 Scenario 10 Obligation of using Class 1 motors in any part of A/C equipment ................................ 35 Scenario 11 Obligation of using the best fans (0.25 W/(m3/h) by characteristic curve adaptation) in any part of A/C equipment................................................................................................................. 35 Scenario 12 Cooling Towers disappear and the areas cooled are retrofitted with air cooled chillers when they become obsolete, but the secondary equipment remains the same ................................... 36 Scenario 13 Obligation of moving from decentralised systems to centralised systems at a certain value of capacity (Portuguese regulation).......................................................................................... 36 Scenarii 14 &15 Reversibility is compulsory in all segments where reversibility is feasible............ 37 General conclusions ........................................................................................................................... 37

8. EFFICIENCY RATING AT PART LOAD: AN IPLV FOR EUROPE.......... 39

8.1 The importance and nature of part-load management measures.............................................. 39 Importance of establishing a EU method about part load .................................................................. 39 How to reduce the capacity of a chiller? ............................................................................................ 39 Staging of Part capacity (control issues) ............................................................................................ 41 High pressure control at part load ...................................................................................................... 42

8.2 Is the IPLV approach directly applicable to European conditions? ......................................... 44 Buildings used in deriving the US-IPLV ........................................................................................... 44 Climate used in IPLV derivation ....................................................................................................... 44 Building cooling load calculation in US-IPLV.................................................................................. 45 Calculating US weighing coefficients................................................................................................ 45 Interpolation scheme needed to reduce testing time .......................................................................... 45 EMPE: an answer to a need for a European weighting with IPLV-like testing ................................. 46 Reduction of EMPE or IPLV to 2 points with extrapolation ............................................................. 47

8.3. Construction of a data base of EU chillers at part load –understanding part load................. 48 Testing conditions and available testing results................................................................................. 48 Impact of load reduction on the efficiency – a reporting format proposed to Eurovent..................... 49 Water cooled chillers –experimental results ...................................................................................... 49 Air cooled chillers –experimental results........................................................................................... 51

8.4 Derivation of a new SEER method (ESEER) .............................................................................. 52 The simulations leading to the reference values of SEER (HSEER) ................................................. 52 Sizing issues for chillers rating as shown by the simulation of the buildings .................................... 52 Reduction of European hourly load curves to a set of four conditions (based on the example of Milano) .............................................................................................................................................. 54 Results for more extreme weather conditions (London, Seville, different distribution systems) ...... 56 Extrapolating to the European stock of chillers in use....................................................................... 58

8.5 Is there a method good enough for classification of products by order of merit?.................... 60 EECCAC final figures -Simplification of the figures and uncertainty estimate ................................ 60 Classification : who is right?.............................................................................................................. 61 EER is a poor selection tool............................................................................................................... 61

IPLV and EMPE are more accurate than EER for classification but do not give enough accuracy for comparison of chillers........................................................................................................................ 62 The proposed ESEER method allows grading and ranking of chillers by order of merit .................. 63 First way to realise the testing needed for the ESEER proposed certification method ...................... 63 Second way to realise the testing needed for the ESEER proposed certification method.................. 66 Final choice of the ESEER testing methodology ............................................................................... 67 Perspective of the proposed ESEER .................................................................................................. 67

9. POLICY OPTIONS AND RECOMMENDATIONS TO IMPROVE CAC ENERGY PERFORMANCE.......................................................................... 69

9.1 Some fundamental considerations regarding policy measures.................................. 69

9.2 Policies and measures to encourage the selection of more efficient equipment........................ 69 Measures to provide information to end-users and equipment procurers .......................................... 69 A to G efficiency grading of central air conditioner components ...................................................... 70 Market mixed statistics based on the scheme (splits and packages mixed) ....................................... 78 Removing less efficient equipment from the market (MEPS and voluntary agreements) ................. 79 Encouraging the selective acquisition of more efficient equipment by other means ......................... 80

9.3 Policies and measures to encourage the adoption of more efficient system structures............ 81 Policy aims and potential measures targeting the adoption of more efficient system structures ....... 81 Legal basis for policy measures targeting more efficient system structures ...................................... 82 Specific recommendations ................................................................................................................. 83

9.4 Policies and measures to improve system maintenance and operation ..................................... 85 Policy aims and potential measures targeting improved O&M.......................................................... 85 Legal basis for policy measures targeting O&M ............................................................................... 85 Broadening the application of existing policy measures addressing O&M ....................................... 86 Specific recommendations ................................................................................................................. 86

Definitions and general terms used in the study ............................................................................... 88

List of abbreviations ............................................................................................................................ 88

REFERENCES..................................................................................................................................... 90

3 See final report to contract no. XVII/4.1031/Z/ XXXX co-ordinated by Prof. Roriz, IST, Lisbon. 4 Tertiary is a European word indicating all human activities and related buildings other than industry or households

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6. TECHNICAL AND ECONOMIC EVALUATION OF THE ELEMENTARY EQUIPMENT USED IN CAC 6.1 Energy-engineering analysis of chillers Chiller prices as a function of the refrigerating fluid and EER Chillers using the refrigerant R407C, which has been developed as a zero ODP substitute to R22, on average have an identical energy performance and do not appear to be any more expensive to purchase, judging from an analysis of their publicly quoted prices. Figure 6.1 shows the price of the equipment as a function of its refrigerating power and refrigerant. From this it appears that there is no additional cost for chiller equipment that uses R407C compared with those which use R22. Figure 6.1. Chiller cost versus cooling capacity, as a function of the refrigerant

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Role of condensing medium By contrast there appears to be a much stronger relationship between chiller price and the choice of condensing medium. The relationship between chiller capacity, condensing medium and price is shown in Figure 6.2. For small capacities, the difference of cost between less expensive water condensation chillers (needing an outside tower) and more expensive (but complete) air condensation becomes small and does not pay for the additional equipment necessary for the system with condensation on water. The use of cooling with water can only be economically justified in large capacity systems.

Figure 6.2. Chiller price vs. cooling capacity as a function of the type of condensing medium

y = 230.98x + 5130.2 R2 = 0.7835

y = 122.46x + 4944.3 R 2 = 0.9228

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Additional costs for reversibility The cost of a reversible chiller is on average 10% higher than the cost of a traditional cooling-only chiller. The data in Table 6.1 show a comparison of prices and EER for a sample of 89 cooling-only models with 44 reversible models.

Table 6.1. Comparison of prices between cooling-only and reversible chillers

reversible cooling onlyCoolingCap. kW

Price(euro)

EER Price(euro)

EER Price differencein %

28 kW 8590 3.5 (water) 7629 4 (water) 12.6%33.5 kW 12498 2.33 (air) 11629 2.33 (air) 7.4%64 kW 18900 2.57 (air) 17025 2.34 (air) 11%

Although the cost of reversible equipment is on average 10% higher than the cost of a cooling-only system, it may well be offset by the avoided installation cost of a stand-alone heating system.

Defining chiller part-load efficiency An exercise done within AICARR (the HVAC engineers association of Italy) showed that the ARI coefficients were completely unsatisfactory for use in Europe and this has lead to a proposal known as EMPE (European Method for Part Load Efficiency). However the ARI standard is very important because it is one step further than the ISO TC 86 / SC6 / WG9 part load testing standard which is being elaborated internationally. Many chillers have been tested under the US IPLV approach and its introduction has produced a significant market transformation impact in the US. Thus an integrated part-load testing approach is no longer a hypothetical proposal, but a practical tool.

As the result of a CEC mandate, TC 113 of CEN is developing a part load test (part capacity in fact) applicable to any AC equipment (CEN02). The difference between part load and part capacity lies in the extent of the testing. In part-load testing the manufacturer manually adjusts the chiller to attain the required testing load which is varied by conditions in the test chamber. In part capacity testing the entire chiller is tested such that the chiller control unit, is used to adjust the cold generated by the chiller to a given percentage of the full output, i.e. the real environmental test conditions are not applied and hence the feedback between the outside and the chiller control is not tested. The draft standard (No. Part5) which it is proposed would be added to the CEN 814 and 255 standards includes a part capacity test with 50% input power and the same temperatures as the full load. This proposal received many negative remarks from the European national standards bodies who vote on the adoption of CEN standards and thus will first be used as an ENV text (to be used on a voluntary basis). A positive interaction took place between CEN TC 113 and EECCAC, allowing a better representation of European interest in those subjects.

Italy has started its own part load testing (Italian Standard UNI 10963 " Air conditioners , chillers and heat pumps- part load tests."). This Italian standard provides more ambitious a starting point than the CEN draft standard resulting from the CEC mandate. Besides full load conditions, one temperature regime is tested first at full load, then with part load capacity. In any case the laboratory shall run at least one test with a capacity from 20% to 30% of the nominal full load capacity. The technician helps the equipment reach this point by reducing the swept volume but if the unit has only on-off control, the test has to be run cyclically. The duration of the cyclic run test is one hour and shall include at least two cycles.This gives an idea of the chiller EER in the most common operating conditions and allows additional test runs to be performed for any load. The manufacturer can run supplementary tests at different part loads, supplying more results in order to

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get more accurate calculations and data. The connection between this testing standard and the EMPE (IPLV like) method seems still to be established.

As part of the EECCAC study, EDF (Electricité de France), Eurovent-Certification and Armines launched an experimental exercise in order to prepare the ground for a wider use of part load testing in Europe. The exercise has two strands. First, some part capacity points will be tested in the EUROVENT certification programme in order to gain some testing experience. EUROVENT could require their current full-load measurements to be supplemented with testing at additional points in order to better characterise the annual behaviour of the equipment. This would occur at a relatively low cost as the chiller is already being installed on the test bench. Second: some chillers (one of which will be a reversible air to water heat pump) will be fully tested (i.e. with full performance mapping) using EDF’s facilities.

Available data and simulation tools Eurovent – Certification runs a directory of products on the EU market which gives good information of product performance. The Directory has been used extensively. We have had access to a number of simulation tools, two from manufacturers & THERMOPTIM, a thermodynamic software from Armines. We have also used EUROVENT testing points (Joint project) and EDF’s experimental testing programme.

Incremental costs as a function of efficiency Our first approach has been to disaggregate the total cost of the chillers into their main components and then extrapolate the cost of each part assuming increasing efficiencies. This method overestimates the costs of efficiency : we know that Energy Efficiency costs less than expected based on such a method when it's taken on board by the companies because then R&D can intervene. But to start the process we need to find the margin for self paying improvements on the market. We have based our analysis on a base line chiller with a screw compressor operating with R134a.

Starting from such values one can seek the level of performance ensuring the minimum of LCC but one can also estimate the overcost associated with some levels of standardised performance (moving from G, to F, to E...) that the market cannot reveal (there is only one market price if we exclude the brand name effects). An extrapolation could also determine the total overcost of the industry and the price increase to be expected from EE improvements. The simulations have been performed with Thermoptim®. This extension of a commercial software enables to perform non nominal performance calculations according to the following description of the components of the modular chiller:

• compressor: isentropic efficiency and volumetric efficiency as a function of the compression ratio,

• evaporator: two zones, biphasic and vapour ; one correlation by zone gives the heat exchange coefficient ; the parameters are only physical, exchange area, free flow area, hydraulic diameter and an intensification factor to take into account specific surface enhancement or increase,

• condenser: 3 zones ; one correlation by zone gives the heat exchange coefficient ; the parameters are only physical, exchange area, free flow area, hydraulic diameter and an intensification factor to take into account specific surface enhancement or increase,

• expansion valve: the expansion process is supposed to take place at constant enthalpy.

The base case correspond to an air to water chiller with a screw compressor, working with the R134a refrigerant, with a Cu-Al air coil and a shell and tube evaporator. The equipment has been designed to represent similar behaviour, in terms variation of the EER with outside air temperature and water temperature, to chillers’ manufacturer whose data were available.

The nominal full load efficiency has been decreased by decreasing the compressor isentropic efficiency and the exchange coefficients at both heat exchangers, in order to represent the “bottom” of the market in terms of performance and so to represent what it would cost to request a minimum performance to all chillers. To complete the market reality, a similar work should be performed with an air to water scroll chiller with R407C as the working fluid. A similar study should be made also on a water cooled chiller.

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Optimisation of the chiller used as baseline without any system consideration Now we can perform some economic calculation and compare the improvements proposed with the diversity on the market. For a given electrical power the capacity varies proportionally to EER; for a given capacity, the compressor can be reduced when EER increases. So the cost per kW decreases with the first steps of performance and only increases later (see figure 6.3).

Figure 6.3 The cost of a chiller at nominal capacity according to its EER

Conclusion : the best chiller having the same cost (100 Euros/kW) as the present “worst” has an EER around 2.80. The range from 2.00 to 2.80 shows reasonable prices for a chiller judged only on capacity. It corresponds exactly to the present market. The minimum cost chiller according to our analysis has the same EER as the average market EER 2.50), which may be considered as a validation of our cost reconstruction.

Optimisation of a chiller in a system The energy consumption of equipment will be more and more considered in the equipment design process. One day, a definition of chillers performance based on SEER and SCOP will be substituted to the ones given as EER and COP. The part load benefits will then be optimised and the optimisation can then be made on the basis of SEER (computed here with the EMPE method). So it is interesting to define the “optimum” taking into account consumption. The search for the optimum has been done in the same way, through successive additions, including part load options (Figure 6.4).

Figure 6.4. The cost of the service rendered by a chiller in terms of SEER

Optimisation of Cost/kW final

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The optimal level of performance for a screw chiller is about 40% more efficient than the present « bottom » of the market. We have seen that the chiller chosen in chapter 5 to represent the stock has a SEER around 2.00. The optimal chiller is already present on the market. It has a SEER between 3.00 and 3.50. One way to reach this performance is an EER around 2.46 (enhanced evaporator and condenser, improved compressor) and a splitting in 3 or 4 scroll units of the capacity of the compressor. We have estimated the associated overcost at 12.3 Euros/kW (+12.3%). Once again, manufacturers engineers may have other ways to reach 3.25 SEER, less expensive, but our objective was to find out if there is a margin for improvement. There is a large margin for improvement and central solutions are not condemned in comparison with packaged units if they improve their performance.

An example can be find in table 6.1 hereunder. For the bottom of the market we have used the same chiller as for the stock : 2.5 EER with a poor part load control (uncontrolled screw). For the best range of products on the market we consider the same nominal EER but the best part load behaviour we found experimentally on a 4 scrolls chiller. Part load optimisation would bring 10 to 20% decrease of the total bill of the office building simulated (SSEER from moving from 1,16 to 1,34, from 0,64 to 0,73, from 1,04 to 1,22). The relative change at that bill level is half of what it is at chiller level due to the weight of auxiliaries remaining unchanged.

Water cooled chillers Another type of improvement could be to use a water cooled chiller (with a better nominal EER). Cooling tower use is desirable in principle because the EER seems better. However there are pressures against CT due to the legionella problem. There is also a water consumption, not to be forgotten. The question is : will Cooling Towers remain used in the present 12% of cooled area, or will they disappear slowly? The promoters of this solution should compare the LCC with and without CT, their water consumption and their cost.

Then we have the use of natural water, with a higher initial cost. The solution expands so slowly that we don’t have figures to prove any growth. Not only EER is better but it is kept most of the year due to the constant natural resource. However there are strong electric auxiliaries for pumping and circulating the water, namely in confined aquifers (the case represented here).

We have simulated for the previously defined office building both water cooled system, all conditions remaining the same (figure 6.2). The water cooled chillers are a screw unit with nominal 3.31 EER for the

Table 6.1 performance of improved air cooled chillers

kWh/m2 for cooling with the CAV system TAC

“poorest” WC chiller with CT

“best” WC chiller with CT

“poorest” WC chiller with natural water

“best” WC chiller with natural water

“poorest” AC chiller

“best” AC chiller

Seville 100,19 84,82 100,75 84,80 99,26 85,97

London 31,78 28,23 32,60 29,37 32,77 28,66

Milan 69,64 59,66 69,66 59,87 70,49 60,15

Table 6.2 performance of improved water cooled chillers

SSEER for cooling with the CAV system TAC

“poorest” WC chiller with CT

“best” WC chiller with CT

“poorest” WC chiller with natural water

“best” WC chiller with natural water

“poorest” AC chiller

“best” AC chiller

Seville 1,15 1,36 1,14 1,36 1,16 1,34

London 0,66 0,74 0,64 0,71 0,64 0,73

Milan 1,05 1,23 1,05 1,22 1,04 1,22

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The energy benefits of water compared with air appear small, once you include all the auxiliaries needed to reach natural water or to run a cooling tower and the climatic differences don’t change that comparison. What is important is the part load behaviour of the chiller not its type!

6.2 Engineering approach of the performance of Packaged units For the most part the large packaged air conditioners, such as ‘roof tops’, found in the European market are either identical to, or share the same technology as, models available for sale in wider international markets, such as the USA. Bearing in mind this technological similarity and the resource constraints applying to the current study a decision was made to adapt the results of existing techno-economic energy engineering analyses conducted for this type of equipment in the USA for use in Europe rather than conduct a fresh European analysis.

The US Department of Energy imposed minimum energy performance requirements for large packaged air conditioners (known as ‘unitary air conditioners’ in the USA) through the EPCA in 1992. As recently as 1999 the non-binding ASHRAE 90.1 standard proposed minimum energy performance requirements for the same appliances and these have since been made mandatory requirements at the state level by a large majority of US states. In 2000 the US DOE launched a revision process for the existing EPCA MEPS which aims to set more stringent MEPS from 200X. Following the US MEPS development process a full techno-economic energy engineering analysis has been conducted for large packaged air conditioners, which forms the basis for the results reported in this section. An analysis of the US market for large packaged central air conditioners established that the market could be adequately represented by a techno-economic energy engineering analysis of two fundamental models: 1) a roof-top unitary air conditioner having a cooling capacity of 7.5 tons (26 kW), and 2) a roof-top unitary air conditioner having a cooling capacity of 15 tons (52 kW).

A parallel analysis of the European market shows that the average cooling capacity of large packaged AC units in the EU is 28.9 kW while that in the USA is 36.2 kW. Figure 6.5 shows the distribution of models by cooling capacity in the two markets. As a result the smaller 26 kW base case unit is much more representative of the type of models found on the EU market than the 52 kW unit.

The only other significant difference in the products found on the two markets concerns the average energy efficiency levels. As a result of the existing US regulations the minimum permissible EER for packaged AC units with a cooling capacity between 19 and 39.5 kW is 2.61 W/W and for those with a cooling capacity between 39.5 and 70.3 kW is 2.41 W/W. In 2003 the lowest efficiency unit which was active on the US market had an EER of 2.5 W/W and the average efficiency was 2.9 W/W. The maximum EER level found on the US market in 2003 was 4 W/W. The lowest EER considered in the US energy engineering analysis is 2.78 W/W for both the 26 kW and 52 kW units. By contrast the average efficiency of packaged units in the EU market and within the Eurovent database was 2.46 W/W in 1998, the minimum EER was 1.78 W/W and the maximum EER was 3.58 W/W. The large difference in the lower and average efficiency levels can be ascribed to the impact of the US policy measures and the absence of equivalent measures in the EU.

Figure 6.5. Share of large packaged air conditioners as a function of cooling capacity in the EU and US markets (source: Eurovent and ARI databases)

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The US energy engineering analysis This section draws heavily from (TIAX 2002).

The goal of the US energy engineering analysis was to develop cost versus efficiency curves of large packaged AC units to guide policy development. In particular the intention was to determine the life cycle cost of packaged AC units as a function of their EER. The methodology used was as follows:

A total of eighteen large packaged AC units, representing several manufacturers and a wide range of efficiency levels were examined and four units chosen. The selected units were broken down (physically or using catalog/design data) to create a bill of materials that was fed into a cost model.

The cost model itemises ‘fixed’ factory expenses such as: equipment and plant depreciation, tooling amortisation, equipment maintenance, utilities, indirect labour, cost of capital and overhead labour and ‘variable’ factory expenses such as: manufactured materials, purchased materials, fabrication labour, assembly labour, shipping and indirect materials. It also itemises corporate expenses such as: research and development, net profits, general & administration costs, warranty costs, taxes and sales and marketing costs.

The inputs to the cost model were reviewed by individual manufacturers and the values adjusted if appropriate. The cost efficiency relationships established in this way were found to follow an exponential growth curve, thus the data for each manufacturer was regressed to a exponential curve.

Each of these curves was in turn regressed to a single market-average exponential curve to give the results shown in Figures 6.6 and 6.7 below. These figures also show the upper and lower 95% confidence intervals as broken dotted lines about the average line (solid).

Figure 6.6. The incremental cost of 26 kWc packaged air conditioners as a function of their efficiency on the US market in 2001 (US$) (source: TIAX 2002)

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Figure 6.7. The incremental cost of 52 kWc packaged air conditioners as a function of their efficiency on the US market in 2001 (US$) (source: TIAX 2002)

In addition to using this market-based reverse engineering approach a classic design option analysis was also conducted to explore the cost efficiency relationships of products with potentially higher efficiency levels than those found on the existing US market .

Life cycle cost analysis This section draws heavily on LBNL 2002.

The data on manufacturing cost as a function of efficiency derived from the energy-engineering analysis were converted into life cycle cost vs. efficiency curves in the following manner: Mark-ups for wholesalers, distributors and mechanical contractors were determined (the latter separately for small and large contractors operating on new- or replacement-construction markets). Incremental changes in the total installation costs as a function of the energy efficiency of the large packaged AC units were estimated by applying the incremental wholesaler, distributor and contractor mark-ups to the incremental ex-factory equipment costs estimated in the energy-engineering analysis.

The results indicate that on average the life cycle cost minimum in the USA occurs for an EER of 11.5 Btu-hr/W (=3.37 W/W) for both the 26 and 52 kW units. Data has been gathered in the EU on: the average efficiency of packaged AC systems; the typical price and installation costs of packaged systems as a function of their cooling capacity; typical building load factors and marginal electricity prices, that are all necessary inputs if the life cycle costs of European packaged systems are to be determined. However, in order to establish the relationship between these life cycle costs and the nominal efficiency of the packaged system it

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EquipmentCostDelta($)

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is necessary to adapt the US cost-efficiency data to reflect the circumstances in the EU in the absence of equivalent European data. The assumptions used in this calculation imply that not only the trend in relative manufacturing cost vs. efficiency for packaged AC units is the same in the EU as in the USA, but also that the relative trends in distribution, installation and maintenance costs are the same. However, when the life cycle cost results produced in this manner were compared with those produced for large packaged systems with an EER of 2.25W/W derived from the values quoted in the tables of section 2.2, the results were found to agree to within 0.3%! This implies that the adapted US equipment cost versus efficiency relationships are reliable for use in the EU.

For 26kWc units the US analysis implied an average equivalent of 2097 hours of full load operation per year while 800 hours per year is deemed more likely for the EU. The results of the analysis taking these factors into account is shown in Figure 6.13 for the 26kWc unit, which is most representative of the EU market. They show that the life cycle cost minimum occurs for large packaged units with an EER of 3.22 W/W when a 6% real discount rate is applied.

The comparable results for the 52kWc unit are shown in Figure 6.14. Although the overall life cycle cost per kW are lower for the 52 kW unit the minimum still occurs for an EER of 3.22 W/W.

Figure 6.13: Estimated average life cycle cost per m2 of cooled space per year vs. EER for large packaged air conditioners on the EU market (based on a 26kWc unit)

€10

€11

€12

€13

€14

€15

€16

€17

2.2 2.4 2.6 2.8 3.0 3.2 3.4 3.6EER (W/W)

ALL

C (€

/m2/

yr

Figure 6.14: Estimated average life cycle cost per m2 of cooled space per year vs. EER for large packaged air conditioners on the EU market (based on a 52kWc unit)

18

€11

€12

€13

€14

€15

€16

€17

2.2 2.4 2.6 2.8 3.0 3.2 3.4 3.6EER (W/W)

ALL

C (€

/m2/

yr

6.3 Energy Efficiency of Air Handling Units seen as tradable goods Major European manufacturers, members of the Eurovent WG 6C “Air Handling Units “ took a very active part in the SAVE project concerning Life Cycle Cost of equipment used in ventilation, air conditioning and refrigeration. The fact is that the energy consumption (operating) costs represent more than 80% of the total cost during the lifetime of the unit. Investment and maintenance together make only about 20% of the Life Cycle Costs. Decreasing the Life Cycle Cost means in fact in the same time an important energy saving.

The Eurovent Working Group prepared a Recommendation for selection and design of an Air Handling Units in order to reduce or minimise Life Cycle Cost. Such a Recommendation was not simple to make, essentially because large differences exist between European countries concerning climate conditions and also energy prices (electricity, energy for heating and cooling). It was emphasised that the most important parameters influencing the Life Cycle Cost are:

• size of internal area

• pressure loss in the duct system

• heat recovery device

• control systems for regulation of the actual demands of the ventilation.

In many cases it may be better to select a larger unit in order to reduce the operating costs. The internal pressure drop will decrease and efficiency of other functioning parts will be better. Even if the investment cost increases with a larger unit, the pay back will be better.

Fans integrated in AHU A well designed duct system with low pressure losses will greatly reduce the electrical consumption of fans. For instance a reduction of pressure loss from 400 to 250 Pa ( which may be relatively easily obtained ) will give much better results than increasing the efficiency of the fan by 5% ( which may be difficult to achieve ). There are two effects in the same direction of the electricity consumption in fans : direct electricity consumption and increase of cooling loads leading to an indirect increase. It is not a small effect : in Seville the system with less fans demands 95 kWh/m2 in cooling as opposed to 105 for the system with more fans; in London the figures become smaller but the difference larger (25% in cooling demand).

Table 6.3 indicates the minimum efficiency levels for fans recommended by Eurovent ( Efficiency figures valid for most common operation hours (around 3000 hours/year). Higher efficiency values are strongly recommended for longer operation periods.

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Table 6.3. Minimum efficiency levels of fans recommended by Eurovent

[m3/s] [m3/h] 250 315 400 500 630 800 1000 1250 1600 2000 2500

1,00 3600 30% 31% 32% 33% 35% 36% 37% 39% 39% 40% 40%

1,25 4500 31% 32% 33% 34% 35% 37% 38% 39% 40% 40% 41%

1,60 5760 32% 33% 34% 35% 36% 38% 39% 40% 41% 42% 42%

2,00 7200 33% 34% 35% 36% 38% 39% 40% 41% 42% 43% 43%

2,50 9000 34% 35% 36% 37% 39% 40% 42% 43% 43% 44% 45%

3,15 11340 35% 36% 38% 39% 40% 42% 43% 44% 45% 45% 46%

4,00 14400 37% 38% 39% 40% 42% 43% 45% 46% 46% 47% 48%

5,00 18000 39% 40% 41% 42% 43% 45% 46% 47% 48% 48% 49%

6,30 22680 40% 41% 42% 44% 45% 46% 48% 49% 50% 50% 51%

8,00 28800 42% 43% 44% 45% 47% 48% 49% 50% 51% 52% 52%

10,00 36000 43% 44% 45% 47% 48% 49% 51% 52% 53% 53% 54%

12,50 45000 44% 45% 47% 48% 49% 51% 52% 53% 54% 54% 55%

16,00 57600 45% 46% 48% 49% 50% 52% 53% 54% 55% 55% 56%

20,00 72000 46% 47% 48% 49% 51% 52% 53% 54% 55% 56% 56%

25,00 90000 47% 48% 49% 50% 51% 53% 54% 55% 56% 56% 57%

MINIMUM TOTAL EFFICIENCY* FAN/MOTOR COMBINATION [%]

Air flow rate Available static fan pressure [Pa] **

Heat recovery section of AHU By using Heat Recovery Systems it is possible to reduce the energy consumption and consequently the Life Cycle Cost tremendously - especially with extreme climate conditions in both, cold or hot climates.

A certain level of heat recovery is recommended by Eurovent, Table 6.4, by taking advantage of the results of the Eurovent LCC study.

Table 6.4. Levels of heat recovery as recommended by Eurovent, depending on the number of hours of operation per year (h/a)

Annual hours of operation (h/a) h/a ≤ 3000 3000 < h/a ≤ 6000 6000 < h/a ≤ 8760

20

Heat recovery wheel - min. dry efficiency* - max. pressure drop

65%

200 Pa

70%

150 Pa

75%

125 Pa Plate heat exchanger without bypass - min. dry efficiency* - max. pressure drop

45%

250 Pa

50%

200 Pa

55%

150 Pa Plate heat exchanger with bypass - min. dry efficiency* - max. pressure drop

40%

300 Pa

45%

250 Pa

50%

200 Pa Heat pipe - min. dry efficiency* - max. pressure drop

45%

300 Pa

50%

250 Pa

55%

200 Pa Run around coil loop - min. dry efficiency* - max. pressure drop

40%

300 Pa

45%

250 Pa

50%

200 Pa * Dry efficiency based on a mass ratio of 1 Running the Air Handling Unit at a speed which is needed for the actual demand will also save energy. Fans using the inverter, give possibility to maintain the optimum speed for different air flow rate demands during the day.

In the Eurovent Recommendation there are many examples for various European conditions. It is also possible to see how different parameters influence the Life Cycle Cost and the consultant or purchaser may look at the special conditions that are valid just for his particular system for the case of a cross flow sensible heat recover device with an efficiency of 0.6. Fan consumption remains constant. Heating saving potential is very important in every climate and it is over 40%. Cooling savings are less important and only significant in hot summer locations (3% in Seville). The combination between heat recovery and free cooling has been proven by simulation to be a simple addition of savings.

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7. TECHNICAL & ECONOMIC EVALUATION OF CAC SYSTEM PERFORMANCE AS A FUNCTION OF THE DESIGN OF THE AC SYSTEM We need to simulate different CAC systems in order to a) understand the relative importance of each aspect of the CAC system in influencing overall CAC energy consumption, b) evaluate the relative energy performance of different systems to enable fair comparisons to be made between systems.

7.1 Comparison of different CAC systems

Energy consumption for a given comfort level The selection of a system takes into account a number of specific factors (customer’s demand, noise, geometry of building) and the comparisons could in principle only take place between 2 or 3 systems, once the building is designed. We can however study the imaginary situation where all systems are feasible and where decision would be made on the basis of energy and cost only. To give an idea of the cost of comfort, we computed here the ALCC of a few systems installed in an hypothetical 2000 m2 building (800 hours equivalent, optimised SSEER).

Table 7.1 Hypothetical ALCC of a few air conditioning systems for a 2000 m2 building with comfort level TAC

Components of cost ALCC

Investment Energy

Euros/15 y Euros/15 y Euros/m2/y

RAC with Primary Air

TAC

Hypothetical

SSEER

Multi Split systems 2,25 248000 128000 20,75

Packaged systems (under windows)

2,25 188000 128000 16,77

CAC - Central Air Conditioners

TAC

Large packages (Roof tops...)

2,25 130000 128000 12,91

Large splits with primary 2,25 274000 128000 22,48

CAV 1,30 336000 221538 34,76

VAV 1,70 352000 169412 34,33

2P FCU 2,00 318000 144000 27,53

4P FCU 2,00 326000 144000 28,10

WLHP 2,40 200000 120000 17,30

VRF with primary air 2,80 348000 102857 31,78 Note: Sizing = 120W/m² for CAC, 240 for RAC

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Table 7.2 Hypothetical ALCC of a few air conditioning systems for a 2000 m2 building with comfort level TC

Components of cost ALCC

Investment Energy

Euros/15 y Euros/15 y Euros/m2/y

RAC without Primary Air

TC

Hypothetical

SSEER

Multi Split systems 2,25 220000 128000 16,69

Packaged systems (under windows)

2,25 160000 128000 13,30

CAC - Central Air Conditioners

TC

Large splits 2,25 186000 128000 14,77

2P FCU 2,00 220000 144000 17,23

4P FCU 2,00 228000 144000 17,68

WLHP 2,40 100000 120000 9,65

VRF without primary air 2,80 260000 102857 22,01 Note: Sizing = 120W/m² for CAC, 240 for RAC

Comparison of costs and sensitivities Note that the solutions not providing total air conditioning but just cooling, cost 5-6 Euros less than the others. The TAC solutions are in the range of 17 to 35 Euros (except the rooftop which is less expensive but has specific geometric constraints and the water loop heat pumps) and that solutions providing TC only are in the range of 15 to 20 Euros (if you are not ready to show packaged RAC in your facade).

The sensitivity to electricity price is significant : –2/+ 3.5 Euros per year and square meter if we consider the extremes of the prices on the EU market. Energy represents 30 to 50% of total expenditures. The values obtained here seem very high compared with what we can read in some places. All investment has been taken as amortised over 15 years which is a heavy assumption, specially if we think of the systems providing ventilation. They have two functions and we could as well decide to make economic calculations on the cooling function alone.

7.2 The improvement of the efficiency of air handling systems in CAC The influence of the main elements of the air handling system have been quantified. The overall efficiency of the air-handling system relates in principle to the efficiency of each of its consituent components, the operating mode, system configuration and operating conditions, the efficiency of air diffusers. in addition the efficiency depends on ducting: specific pressure drop rating, the air tightness rating and the influence of thermal insulation. Fan specific consumption depends not only on fan efficiency, but also on pressure drop in the distribution system. This doesn’t get the prominence that it deserves: because fan energy is proportional to the square of the velocity and velocity is (inversely) proportional to the square of duct diameter, duct sizing can be very important (and of course, filter selection). Moreover the importance of full and part-load operation was stressed and led us to detailed computer simulation (DOE software).

Primary Air and ventilation Two philosophies of ventilation seem to exist in Europe also responding to local natural conditions : in the first one (adopted by Northern countries), ventilation comes first as an hygienic necessity and then a further decision leads to cool the space or not. In the second one (apparently Southern States), the decision of A/C

23

comes first and leads to more air changes with the outside, and to controlled ventilation. As a result Local extraction (LE) is the dominant feature in some countries, while V (central ventilation) dominates others. Obviously V allows a better air quality (dust, temperature, etc.) but is more costly. The energy impact of the two philosophies is large but not really part of our study, but the capacity of heat recovery being very different in the two situations, we have to consider them. Central ventilation (V) has been the base of our study.

There are also in some countries obligations of ventilation in cascade, the exhaust taking place in the “very polluted” rooms. We have not investigated further this option.

Heat recovery on primary air All what has been said about the improvement of all-air systems and of the heat recovery section of AHU remains true for the primary air of mixed systems. The flow rate is lower and you have to create the exhaust air circulation to the heat recovery station at a cost.

Motors and fans efficiency Overall Fans efficiency can be treated with a power/flow ratio like 0.25 W/(m3/h) for improved systems against 0.75 W/(m3/h) for the worst. Note that W/(m3/h) is a measure of Delta P/efficiency; it includes the fan as a component and the design effort on the air circuit. This is the normative parameter in the USA. You cannot install air distribution systems if SPF is not under a certain value : 0.47 W/(m3/h) for CAV (SPF = 1.7 W/(l/s)) and 0.57 W/(m3/h) for VAV (SPF = 2.05 W/(l/s)). A similar rule exists in the USA for the ventilation aspect of rooftops, etc. More speculatively, parameters like the power/flow ratio or the combined efficiencies of motors/fans used by Eurovent in AHU should - in principle - be applied to the whole air supply (and extract) system in a proper building thermal regulation.

For existing CAV system type, we used an average specific consumption of 0.47 W/m³/h. In fact, we assume that a 15% reduction of SC (0.4 W/m³/h) may be achieved by the use of high efficiency fans with an overcost of 2 Euros/fan kW. Conclusions of DOE simulations: logically, fan consumption suffers a 15% reduction but also there is a decrease in cooling due to fan heat that ranges from 4 to 12.5% and that we have taken as 8% on average.

Variable air flow and lower head losses In our study, the classic constant flow all-air system serving as a basis and has been assumed to represent the full EU STOCK of Air only systems (34% of CAC), even if its market share is now declining. The direct comparison with DOE simulation of the effects in Europe of variable flow (option called VAV, Variable Air Volume, in American English) has been possible. However fans are different in the two systems, as in the US regulations (CAV=0.47 W/(m3/h) and VAV=0.57 W(/m3/h). The reason why the consumption figure is higher for VAV than for CAV is due to higher head losses in the air distribution network mainly caused by higher air velocities and VAV terminal boxes. In this section we study the sole active unit of HVAC air-side. As previously commented, every system type but VAV is equipped with constant air volume fans. Efficiency index is specific consumption and energy use will be expressed in kWh/m².

This end-use consumption is almost constant along the year for CAV fans, see figure 7.1. The only difference is due to the number of working days per month. This is not the case for VAV systems where differences are due to flow regulation. As an example we represent monthly fan consumption for CAV and VAV system types.

Figure 7.1 Fans energy over months in a CAV and in a VAV system

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CAV and VAV fan consumption for SEVILLE

0

1

2

3

4

5

kWh/

m²

CAV 4,39 3,79 4,39 4,19 4,39 4,19 4,19 4,59 3,79 4,39 3,99 3,99

VAV 0,57 0,5 0,54 0,53 0,66 0,98 1,63 1,93 1,11 0,78 0,49 0,5

JAN FEB MAR APR MAY JUN JUL AUG SEP OCT NOV DEC

VAV consumption reduction percentage is higher during winter because of lower loads, VAV boxes operate at minimum flow, which corresponds to outdoor ventilation air ratio. During the summer there is also a large saving potential that is always higher than 60 %.

Conclusions: VAV saving percentage varies from 75 to 80 % despite specific consumption for fans (0.57 W/(m³/h)) is higher than that for CAV (0.47 W(/m³/h)). The main reason for this large saving potential is the adaptation of zone flow rate to real load conditions which are normally under design values due to load calculation oversizing. On average we could say that VAV save 80% of fan energy in Europe. There are discussions about the reality of those results on the field due to O&M issues : balancing, controllers tuning, etc. Conservatively, we made calculations with a 50% saving value.

The Ashrae approach requests variable speed in all air ducts not only blowing through a target of consumption to be reached at full load and another one at a specified part load value. Europe should take such a measure when applying its “Energy Performance of Buildings” new directive.

Dual Duct final distribution systems (rather inefficient by principle) are only allowed with variable flow in some countries and this measure could be extended to all countries. It seems not to present any specific potential for our study because they are very uncommon.

Terminal reheat issues On air-side it’s also possible to provide evidence (simulation or literature) about the order of magnitude of terminal reheat ; some think it should be completely banned in regulations (except when provided from some renewable source of energy like condenser heat).

In fact it's a promising option to have that perimetral heating or terminal reheat in Air systems (or part of the reheat in an AHU) made from heat recovery on condenser and not from a boiler. The cost and performance of heat recovery from condenser being known, this option could enter the C/B-analysis.

Air Side Free Cooling (Economiser) Air side Free Cooling (FC) is the key option in a cost benefit analysis : small cost, large potential. Free cooling simulations in DOE have been complemented with cost functions. The present use of this feature in EU systems is far from 100%. Many regulations make it more or less compulsory (USA, Portugal). It is related also with the choice of an optimal blowing temperature : the higher , the better for EER and for FC.

Of course, free cooling is very climate and control dependent. We are using by default the classic temperature control. There is a very limited overcost of 2% of overall system costs (ducting, new connections, control) for a reduction of chiller electricity by 20% (no auxiliaries increase). The potential given by simulation is far higher but we have taken this conservative assumption due to potential O&M uncertainties.

25

The real figures given by DOE show that the cooling saving potential is strongly dependent on climate. Oceanic weathers like London offer a more than 80 % reduction of cooling consumption. In any case, and even for very hot climates like Seville, saving potential is over 20%. Pumping consumption is also reduced due to cooling load decrease and pumps "on demand" control.

Quality of Air Diffusion We have in the Italian UNI standard a set of interesting figures, but no cost data. The option of “displacement” allows to lower significantly the loads. In reality a large portion of the loads doesn’t impact comfort because they are too high in the room. For this reason by introducing cool air at low speed at the bottom of the room and leaving it move upwards when there is a heat source, we have a very energy efficient treatment of demand. Research has not yet produced perfectly consistent values for this approach which already works in practice (a few percent of new installations). Displacement is now reported to be less costly by 15% in installation costs but, as in the case of chilled ceilings, the system is unable to answer to the total heating needs in winter and an overcost appears somewhere else.

AHU improvement We should promote the investigation on the development of low pressure drop AHU components (filters, coils, heat exchangers, sound traps, etc.) since the larger part of the fan pressure is dissipated in the AHU.

The size of the AHU should be determined after a LCC analysis considering the annual operating hours, the fan energy consumption and the unit cost of the options for the air speed in the unit (2.0 m/s, 2.5 m/s, 3.0 m/s and 3.5 m/s),

The quality of the duct system should be controlled in some way, for example;

• Mandatory air leakage tests (air leakages of more than 20% are common),

• Use of low pressure drop connection pieces,

• Air balance criteria and equipment,

• Velocity or pressure drops limits,

The piping systems should be designed so that all AHUs should not have balance valves and flow control valves. Every AHU should be equipped with a small variable speed pump that delivered the correct water flow, at each moment (the control signal that normally goes to the control valve would go to the variable speed pump). This measure eliminates the energy pressure dissipated in the balance valves and the flow control valves.

7.3 Other cost & efficiency trade-offs

Water-side efficiency by sizing and control Oversized FCU of classic type will allow higher operating temperatures at chiller’s level hence an improved EER. For instance moving from the classic 7/12 °C regime to 8/13 °C will increase average from 9.5 °C to 10.5 °C. It will at the same time reduce temperature difference between room and FCU from 12.5 K to 11.5 K (with 22 °C inside) and increase requested area. This process can be extrapolated by a few more degrees but two phenomena appear : the exponential nature of heat exchange and an additional demand for blowing power inside the FCU.

Radiant panels or beams are one step further in the same direction. They allow an increase in distribution temperature by using large areas for heat exchange. How to consider the option of radiant panels? since the benefit claimed comes from the change of temperature regime in the water distribution equipment, costs and benefits have been evaluated by extrapolating the temperature effect to the new data both in summer and winter. There is a second order effect coming from the fact that part of the cooling is radiant, entering in a different manner the equation of human comfort. There is also the possibility to operate them directly from a

26

cooling tower where cold is generated by evaporation. We have investigated only the first phenomenon (temperature regime) and only in one case.

In a special configuration of the ceiling panels, they are mounted under a plenum with injected air, likely for increasing the relatively small cooling capacity. If the cooling capacity of the primary air and the capacity of the chilled ceiling are enough together to ‘treat’ the load, the cost seems the same as the one with FCU. In fact the difference appears in winter because the ceiling looses part of its heating capacity compared with a regular FCU and another system is needed in addition.

In another realisation of the radiant system, recently patented by the firm Van Holsteijn en Kemna (VHK), a local unit (having the appearance and the cost of a local radiator or FCU of good quality) combines ventilation (controlled room by room) and radiant cooling. When radiant cooling is not enough, ventilation is started. Entering water temperatures under room temperatures by only 10K would be enough to deliver 200 W/m2 of radiator. This supposes a decrease in Chiller consumption and an increase in Pump consumption. Fans situation and performance is quite unclear and we decided to wait for the next step of development of the system to include it in the analysis.

We have simulated with DOE software this option. The assumption was to determine performance improvement with a small ∆T difference in a FCU4P system (8/13°C instead of 7/12), since the relative returns will diminish if we move further. The electricity saving is very low (0.1 to 0.6%) and we abandoned this solution in the study.

Design of flow in water circulation Efficiency of pumps for water circulation is the first obvious issue. Pump efficiency is the result of motor and mechanical effects, and present average values are 0.8 and 0.62, respectively. If we improve pumps efficiencies, we will save pumping energy. High performance values are 0.85 and 0.67 for motor and mechanical efficiencies, respectively. Pumping saving potential of this measure is 13%. Total saving potential is low (from 0.7 to 1.7%) due to the limited importance of pumping energy.

Another philosophy (advocated by one company) explains that for the same average fluid temperature in the final FCU (consequently its cooling capacity) one should look for the largest possible DeltaT between inlet and outlet. On the examples taken from company documents, there is a significant benefit at system level from doing this.

The piping systems could also be designed so that all equipment should not have balance valves and flow control valves. Every equipment should be equipped with a small variable speed pump that delivered the correct water flow, at each moment (the control signal that normally goes to the control valve would go to the variable speed pump). This measure eliminates the energy pressure dissipated in the balance valves and the flow control valves.

Our base case assumption is that every system has three-way valves and for this reason circulation loop and pump flow are constant. In a few DOE simulations, we valued the saving potential of using two-way valves and variable flow pumps, that is, pump is controlled by a variable frequency drive. Decrease in cooling consumption due to pump heat ranges from 3.6 to 5.1% (less demand since the heat is not dissipated by the pumps). High pumping savings range from 60 to 72% (relative to pumps consumption).

The inversion from cooling function to heating function and vice versa is a very general problem of systems and should be manually operated or automated carefully. It may have a low energy penalty if done properly : this is a behavioural more than technical issue, difficult to regulate because very dependent on experience gained with a specific building and climate. In principle it could be simulated by generating extreme behaviours, but we decided this was too uncertain and we abandoned the idea of simulating or regulating this type of intervention.

The modification of the temperature regimes on the chilled water loop (set point and point of application of the set point : departure or return) do change consumption. Even if temperature cannot be increased at design conditions, it’s a good practice to “reset” it off design. Two types of information are available : load and

27

outdoors temperature. The electricity saving is very low (0.1 to 1.3%) and we abandoned this solution in the study.

Influence of terminal equipment Improvement of flow efficiency in FCU and AHU water circuits is possible in various ways : Variable flow pump, better Pump efficiencies, adequate Pump performance curve. A remarkable option at system level is to go for “local pumping systems”. Terminal units demand what they need and the evaporator ‘sees’ a variable flow. Evaporators can accept flow rates decreased by 40% typically. The gains in pumping cost become significant.

We have considered the basic Air and Water System types (Two pipes fan-coil, Four pipes fan-coil, Water loop heat pump). Using four pipes FCU or using the two pipes system based on “change over” will allow to feed the four pipes FCU from heat recovered from the condenser, one of the reversibility approaches.

The importance of additional consumption generated by the 2 pipes FCU with electrical heating versus the 4 pipes –if we want to insure the same temperature all the time- can be estimated. Note that we can estimate that on the market those two solutions are still frequent (2PE was 25% a decade ago and seems now at 10% only; 4P is around 15%). Also we suspect that the electrical resistance is used for main space heating not just for this adjustment…..so it’s the main enemy of reversibility.

Simultaneous demand of heating and cooling Some systems have a capacity to transfer heat from one zone to another. Such advanced multizone systems can be justified by its actual benefits. We have gathered some elements on advanced multizone strategies (WLHP, TWL, VRF,…). The system using RAC on a water loop (WLHP) is relatively frequent (1.5 % of total cooled area) and presents specific energy conservation features : transfer from one zone to another, high EER and COP year round, etc. It seems a relatively frequent solution in commercial malls because it allows individual metering of consumption by each user.

In the same way, the uncommon TWL (a promising two water loops system experimented in France and in the UK allowing simultaneous heating and cooling) can provide simultaneous heating and cooling. Finally VRF is one step further in the same direction. It is one way of operating at variable speed (see part on packaged systems). But it is also an interesting system for transfer between zones demanding heat and cold (but this not always realised).

The DOE simulation allowed us to understand the real order of magnitude of simultaneous heating and cooling. In this relatively complex office building where internal and external zones are treated separately, where various facades receive differently the sun, the effect corresponds to only a few percent of the demand. More precisely, we have computed for each hour with simultaneity the lowest of the two quantities : cooling demand, heating demand and expressed it in percent of demand, either cooling or heating, table 7.3.

Table 7.3 Importance of simultaneous heating and cooling

In % of heating CAC system LO MI SE CAV -0.26% -0.23% -4.90%VAV -0.92% -0.92% -3.03%FC2P -6.24% -5.89% -17.61%FC4P -7.45% -6.16% -23.62%PACK -5.33% -4.67% -18.17%WLHP -6.19% -5.47% -21.49% In % of cooling CAC system LO MI SE CAV 0.59% 0.15% 0.21%VAV 5.91% 1.24% 0.50%FC2P 26.88% 7.42% 3.62%FC4P 29.66% 7.73% 4.78%

28

PACK 32.22% 7.41% 4.08%WLHP 31.06% 7.74% 4.53%

The economics of the transfer are as favourable as expected. The homogeneity of figures between air systems on one hand and all other systems on the other hand is interesting.

Heat rejection Cooling tower fans should have variable speed drives and, in systems with more than one cooling tower, all cooling towers should work simultaneously at all times (this strategy reduces drastically the cooling tower fan energy consumption) Temperature control should be modified.

Using natural water –river, ground water, etc.- as a heat rejection medium is very beneficial in energy terms because the high heat exchange coefficients and low temperature at the condenser improve EER. In some circumstances the chiller becomes useless and the natural water can cool directly the building (see system TWL as an example). Control is easy since underground temperatures are constant. Costs and administrative problems are reported as enormous in Italy and Spain while France maintains such a policy (Aquapac). There is also the possibility to generate DHW (Domestic Hot Water) at a small cost from condensing heat.

7.4 The possible strength of regulatory efforts and the minimum LCC solutions

Concentration of efforts on Air based systems We have concentrated our efforts on the air system which show presently (under the CAV form) the most consumption and the highest cost. The designers need the whole range of solution to cover the domain of geometries and air quality requirements. So the bottleneck to the expression of a global reduction in consumption will be the point (shown hereunder by an array) where the improved air based solutions start not to pay for themselves : the designers will find it is too heavy a constraint.

Figure7.1 Bottleneck of air systems

SPECIFIC CONSUMPTION

ALCC Euros/m2/year

kWh/m2/YEAR

airwaterpackagesrac

The result of optimisation

Table 7.2 OPTIMISATION OF AIR BASED IN SEVILLE

29

For

cooling and

pumps (MWh)

For secondary

fans (MWh)

Total electricity

Initial Cost ALCC of

system (0,10

E/kWh)

ALCC of

system (0,06

E/kWh)

ALCC of

system (0,17

E/kWh) 0 – CAV 121,67 37,76 159,43 1 053 640 30,89 29,84 32,73 1 – VAV-50%/ +6 E/m2

121,67 18,88 140,55 1 090 090 31,56 30,63 33,17

2 – FC –20% /+2 Euros/m2

97,34 37,76 135,10 1 074 710 31,05 30,16 32,61

3 – Fans –8% ;–15 %; +2 E/kW

111,94 32,10 144,03 1 054 140 30,65 29,70 32,31

4 – HR 60% -200Pa; -3% +6 E/m2

118,02 37,76 155,78 1 090 090 31,81 30,78 33,60

5 –Lower HL in AHU –7%/+12E/m2

121,67 35,12 156,79 1 126 540 32,80 31,77 34,61

6 –Optimised chiller – see chapter 6

73,00 37,76 110,76 1 062 390 30,32 29,59 31,60

After sorting and combinations, the optimal trajectory of improvement is given in figure 7.2.

Figure 7.2 Optimising with 6, 10 and 17 cEuro/kWh a full all air system

The optimum if very flat, specially if we get interested with the highest cost electricity. The regulatory measure could be taken anywhere between a 0% and a 60% reduction without generating overcosts (in the LCC definition) in Seville. The optimal regulation would request a 50% cut in electricity consumption compared with a standard CAV design.

7.5 The possible effects of technical scenarios In this section, technology scenarios which impact energy consumption and CO2 emissions will be produced and the results compared with the base case for period of 1990 to 2020. The scenarios are defined by impact time and can be translated at that time in new specific consumptions for the market after that time.

25

27

29

31

33

35

37

39

0,00% 20,00% 40,00% 60,00% 80,00% 100,00% 120,00%% of reference

Euro

s/m

2 ALCC17ALCC10ALCC6

30

Scenario 1 MOVING ALL COOL GENERATORS TO AVERAGE PERFORMANCE All packages RAC and chillers presently under average reach by 2005 the EER level corresponding to the average of present market but part load is not taken into account in Eurovent grading and so the corresponding improvement is not obtained. The policy measure associated is banning some classes of equipment either directly (Directive ) or by voluntary agreement. We can also expect that a certain number of years of labelling and communication by energy agencies reaches the same point, nobody wanting to buy a « poor » image equipment. Air Cooled Chillers : the average being 2.50, the classes E F and G should be banned, the weighted gain is 0.23 on full market average and the factor is 96.2% to be applied to compressor consumption. There will be a consequence for reversible winter heating, that will be provisionally taken as the same factor. Water cooled Chillers : the average being 3.85, the classes E F and G should be banned, the weighted gain is 0.175 on full market average and the factor is 93.0% to be applied to compressor consumption. There will be a consequence for reversible winter heating, that will be provisionally taken as the same factor. Packages and large splits the average being 2.46, the classes E F and G should be banned, the weighted gain is 0.05 on full market average and the factor is 98.0% to be applied to compressor consumption. There will be a positive consequence for reversible winter heating of the measures taken for Summer AC, that has been provisionally taken as the same factor. RAC classes E F and G should be banned and the average gain corresponding is 0.10 on an EER around 2.50 so the factor is around 96% There will be a consequence for reversible winter heating, that will be provisionally taken as the same factor. The effects are given by table 7.3. Table 7.3. Effects of technical scenario 1 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 547 625 674 688 BE 50 117 274 421 552 667 690 DE 25 43 71 122 178 228 255 FI 93 139 206 209 227 238 240 FR 2606 3596 5010 8184 10801 12943 13654 GE 613 1187 2286 3999 5469 6644 7210 GR 674 1399 2909 5345 7161 9178 9428 IR 77 99 127 179 220 247 258 IT 8539 11242 16209 24250 27059 29087 29941 LU 3 5 11 18 23 26 28 NE 206 358 605 688 789 853 871 PO 436 669 1020 2042 3028 3949 4483 SP 7370 11744 19689 28229 33082 37782 38648 SW 227 309 391 377 399 414 416 UK 1443 1861 2359 3217 3781 4158 4289 Total 22660 33141 51636 77828 93392 107090 111097 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 2 THE BEST CHOICE OF COOL GENERATORS FOR THE CUSTOMER BASED ON FULL LOAD INFO On average packages and chillers reach in 2005 the EER level corresponding to the minimum LCC (BAT with present information) but part load is not taken into account in Eurovent grading and so the corresponding improvement is not obtained. The policy measure associated is banning many classes of equipment or a negotiated agreement on average full load performance like ACEA agreement for cars.

31

Chillers Air Cooled : either the average moves from 2.50 (present market) to 2.80 or the classes D E F and G are For water cooled chillers, the classes D, E, F and G are also banned. The average is going from 3.85 to 4.45. It means a 14% decrease in consumption, also applied for reverse heating. Packages and large splits either the average moves from 2.46 (present market) to 3.22 by a voluntary agreement or the classes B to G are banned, the gain is 31.3% on full market . There will be a consequence for reversible winter heating, that will be provisionally taken as the same factor. RAC either the average moves from 2.50 (present market) to 3.20 by a voluntary agreement or the classes B to G are banned, the gain is 0.7 on full market average so the factor is around 75% There will be a consequence for reversible winter heating, that will be provisionally taken as the same factor. The effects are given by table 7.4. Table 7.4. Effects of technical scenario 2 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 547 625 674 688 BE 50 117 274 421 552 667 690 DE 25 43 71 122 178 228 255 FI 93 139 206 209 227 238 240 FR 2606 3596 5010 8120 10458 12275 12711 GE 613 1187 2286 3999 5469 6644 7210 GR 674 1399 2909 5278 6801 8444 8397 IR 77 99 127 179 220 247 258 IT 8539 11242 16209 23924 25635 26484 26399 LU 3 5 11 18 23 26 28 NE 206 358 605 688 789 853 871 PO 436 669 1020 2014 2854 3594 3938 SP 7370 11744 19689 27823 31186 34186 33749 SW 227 309 391 377 399 414 416 UK 1443 1861 2359 3217 3781 4158 4289 Total 22660 33141 51636 76936 89196 99134 100137 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 3 BAT- THE BEST CONSUMER CHOICE WITH PROPER PART LOAD INFO All packages and chillers reach in 2005 the SEER level with the minimum LCC (BAT with upcoming information given by part load testing). Part load is taken into account in Eurovent grading and so the corresponding improvement is obtained. The policy measure associated is banning many classes of equipment or a negociated agreement on average part load performance like ACEA agreement for cars. It happens that only the best equipment of today meets the minimum LCC target. Since chillers performance has been calculated for all systems that contain a chiller with a poor part load behaviour chiller (typical screw single circuit unit) and a good one (typical four step scroll), we have an exact DOE calculation of the benefits, that we extrapolate to the scenario. The chiller nominal EER remains the present average. Only part load behaviour is modified. For other systems Scenario 2 is maintained. The effects are given by table 7.5. Table 7.5. Effects of technical scenario 3 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 540 592 614 608 BE 50 117 274 416 523 611 612 DE 25 43 71 121 170 211 229 FI 93 139 206 207 218 222 219

32

FR 2606 3596 5010 8065 10161 11698 11903 GE 613 1187 2286 3944 5163 6043 6341 GR 674 1399 2909 5259 6695 8229 8101 IR 77 99 127 177 208 226 228 IT 8539 11242 16209 23842 25268 25816 25519 LU 3 5 11 17 22 24 25 NE 206 358 605 680 752 787 782 PO 436 669 1020 2007 2816 3515 3816 SP 7370 11744 19689 27751 30839 33526 32868 SW 227 309 391 374 383 386 379 UK 1443 1861 2359 3175 3583 3792 3795 Total 22660 33141 51636 76575 87392 95699 95425 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 4 GENERALISED FREE COOLING Obligation of introducing free cooling on air side of air based distribution systems starting from a low value of flow rate (Portuguese regulation and Ashrae) There is a reduction in cooling demand which is climate dependant. The new consumption ratios were obtained from DOE2 simulations by climate. Ratios for CAV are applied for all systems.The main gain comes from the cooling demand decrease. The effects are given by table 7.6. Table 7.6. Effects of technical scenario 4 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 529 543 525 491 BE 50 117 274 408 481 529 500 DE 25 43 71 118 155 180 183 FI 93 139 206 201 193 179 163 FR 2606 3596 5010 8070 10181 11740 11978 GE 613 1187 2286 3878 4787 5308 5289 GR 674 1399 2909 5321 7028 8908 9059 IR 77 99 127 174 193 197 190 IT 8539 11242 16209 24150 26607 28271 28882 LU 3 5 11 17 20 21 20 NE 206 358 605 665 682 659 615 PO 436 669 1020 2035 2983 3858 4343 SP 7370 11744 19689 28243 33141 37898 38820 SW 227 309 391 363 341 311 282 UK 1443 1861 2359 3121 3328 3325 3178 Total 22660 33141 51636 77295 90664 101911 103994 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 5 Generalised Heat recovery Obligation of introducing heat recovery on exhaust air at a certain value of capacity (Portuguese regulation) There is a reduction in heating and cooling demand which is climate dependant. However the head losses increase significantly with our calculation when there is full Air Conditioning equipment in place. As a result of our calculation, DOE2 results are not positive results for Summer. There is a benefit of Heat Recovery in Winter only, whichever the heating system. Heat Recovery is a positive option to be recommended when it is associated to the ventilation system in Winter and can be positively substituted by free cooling in Summer.

Scenario 6 British regulation on AC – heating, cooling and air movement- adapted for each EU climate

33

Introduction of a MEPS on total electricity used for Heating ventilating and AC in kWh/ m2; to know the cost we have to evaluate the less costly options, which may be on either side, primary or secondary; national values are different and have been derived from UK with corrections for DD: Table 7.7 Extrapolation of UK regulation kWh/m² Present average New buildings New

installations Austria 225,0 183,6 202,4 Belgium 200,2 163,4 180,1 Denmark 224,9 183,5 202,3 Finland 253,7 207,0 228,2 France 198,1 161,6 178,2 Germany 225,0 183,6 202,4 Greece 201,6 164,5 181,3 Ireland 182,4 148,8 164,1 Italy 201,6 164,5 181,4 Luxembourg 200,2 163,4 180,1 Netherlands 196,0 159,9 176,4 Portugal 201,6 164,5 181,4 Spain 165,3 134,9 148,7 Sweden 253,7 207,0 228,2 UK 182,4 148,8 164,1 The impact has been calculated with the assumption of a weighted mix of both situations : new buildings and new installations. The overall reduction being a 12% reduction, we apply then -12% to each item of AC (fans, pumps etc). The policy instrument would be a clear and harmonised implementation of EPB directive. The less expensive way of attaining the objective is the improvement of chillers. Starting from their present averages of EER and SEER, this policy induces almost no extra cost for any stakeholder, and absolutely no cost provided it’s applied to all manufacturers (and so that they all pass on the costs to the customer). To obtain this “free” market transformation a prescriptive minimum should be applied to local manufacturers and importers at the same time. The effects are given by table 7.8. Table 7.8. Effects of technical scenario 6 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 543 603 636 637 BE 50 117 274 417 532 629 637 DE 25 43 71 121 172 214 234 FI 93 139 206 207 218 224 221 FR 2606 3596 5010 8117 10438 12237 12665 GE 613 1187 2286 3965 5280 6272 6674 GR 674 1399 2909 5302 6926 8700 8762 IR 77 99 127 178 212 233 238 IT 8539 11242 16209 24052 26185 27493 27805 LU 3 5 11 17 22 25 26 NE 206 358 605 682 761 801 803 PO 436 669 1020 2025 2925 3739 4160 SP 7370 11744 19689 28002 32019 35771 35929 SW 227 309 391 374 384 389 383 UK 1443 1861 2359 3189 3649 3914 3961 Total 22660 33141 51636 77192 90325 101277 103135 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579

34

Scenario 7 Portuguese regulation on AC – heating, cooling and air movement- adapted for each EU climate A set of figures have been calculated from the text of the regulation which is more a set of prescriptions than an energy target. The effects are given by table 7.9. Table 7.9. Effects of technical scenario 7 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 527 533 507 467 BE 50 117 274 406 472 511 475 DE 25 43 71 118 152 174 174 FI 93 139 206 200 190 173 155 FR 2606 3596 5010 8029 9962 11313 11380 GE 613 1187 2286 3861 4693 5126 5025 GR 674 1399 2909 5294 6878 8601 8632 IR 77 99 127 173 189 190 180 IT 8539 11242 16209 24024 26049 27252 27513 LU 3 5 11 17 20 20 19 NE 206 358 605 662 670 636 584 PO 436 669 1020 2024 2917 3723 4136 SP 7370 11744 19689 28108 32509 36699 37194 SW 227 309 391 362 335 300 268 UK 1443 1861 2359 3108 3265 3208 3019 Total 22660 33141 51636 76914 88832 98434 99222 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 8 US regulation on AC – heating, cooling and air movement- adapted for each EU climate A set of representative figures has been calculated from this very comprehensive regulation. The effects are given by table 7.10. Table 7.10. Effects of technical scenario 8 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 517 488 426 359 BE 50 117 274 399 432 433 367 DE 25 43 71 116 139 149 137 FI 93 139 206 197 176 148 123 FR 2606 3596 5010 7851 9003 9447 8758 GE 613 1187 2286 3787 4279 4313 3849 GR 674 1399 2909 5163 6162 7142 6597 IR 77 99 127 170 173 160 139 IT 8539 11242 16209 23414 23356 22336 20890 LU 3 5 11 17 18 17 15 NE 206 358 605 651 617 540 456 PO 436 669 1020 1973 2601 3075 3140 SP 7370 11744 19689 27364 29014 30065 28208 SW 227 309 391 356 311 257 212 UK 1443 1861 2359 3049 2989 2698 2330 Total 22660 33141 51636 75023 79758 81207 75581 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 9 Obligation of using Class 2 motors in any part of A/C equipment

35

The corresponding consumption cuts were obtained from previous SAVE studies and applied to our stock. The effects are given by table 7.11. Table 7.11. Effects of technical scenario 9 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 548 627 678 693 BE 50 117 274 421 553 670 694 DE 25 43 71 122 178 228 255 FI 93 139 206 209 227 239 241 FR 2606 3596 5010 8193 10851 13040 13790 GE 613 1187 2286 4002 5489 6683 7267 GR 674 1399 2909 5352 7201 9259 9540 IR 77 99 127 180 220 248 259 IT 8539 11242 16209 24280 27193 29334 30273 LU 3 5 11 18 23 27 28 NE 206 358 605 689 790 855 874 PO 436 669 1020 2045 3042 3979 4529 SP 7370 11744 19689 28266 33262 38130 39118 SW 227 309 391 377 399 415 416 UK 1443 1861 2359 3219 3791 4175 4313 Total 22660 33141 51636 77921 93846 107960 112290 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 10 Obligation of using Class 1 motors in any part of A/C equipment Same method as scenario 9. The effects are given by table 7.12. Table 7.12. Effects of technical scenario 10 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 546 618 662 672 BE 50 117 274 420 545 655 673 DE 25 43 71 122 176 223 247 FI 93 139 206 208 224 233 233 FR 2606 3596 5010 8165 10696 12739 13368 GE 613 1187 2286 3988 5411 6529 7044 GR 674 1399 2909 5333 7098 9049 9248 IR 77 99 127 179 217 242 251 IT 8539 11242 16209 24194 26815 28644 29348 LU 3 5 11 18 22 26 27 NE 206 358 605 686 779 835 847 PO 436 669 1020 2037 2998 3889 4391 SP 7370 11744 19689 28167 32796 37245 37922 SW 227 309 391 376 393 405 404 UK 1443 1861 2359 3208 3738 4077 4181 Total 22660 33141 51636 77648 92526 105454 108857 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 11 Obligation of using the best fans (0.25 W/(m3/h) by characteristic curve adaptation) in any part of A/C equipment The chillers and boilers consumptions are not touched. The fan unitary consumption ratios from the best standards are applied to all fans. The effects are given by table 7.13.

36

Table 7.13. Effects of technical scenario 11 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 544 610 647 652 BE 50 117 274 419 539 642 654 DE 25 43 71 121 173 218 240 FI 93 139 206 207 219 225 223 FR 2606 3596 5010 8166 10703 12752 13390 GE 613 1187 2286 3978 5350 6411 6876 GR 674 1399 2909 5345 7160 9177 9429 IR 77 99 127 178 215 238 245 IT 8539 11242 16209 24252 27067 29107 29983 LU 3 5 11 17 22 25 27 NE 206 358 605 684 767 813 819 PO 436 669 1020 2043 3032 3957 4496 SP 7370 11744 19689 28265 33250 38104 39095 SW 227 309 391 374 386 392 387 UK 1443 1861 2359 3200 3698 4005 4086 Total 22660 33141 51636 77794 93191 106714 110603 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 12 Cooling Towers disappear and the areas cooled are retrofitted with air cooled chillers when they become obsolete, but the secondary equipment remains the same Retrofit new water cooled systems by air cooled systems from 2005 on. The effects are given by table 7.14. They prove that this technical transformation can be made without significant energy impact. Note that the cost of the measure would be huge. Table 7.14. Effects of technical scenario 12 on the electricity consumption of countries and EU15 in MWh Country 1990 1995 2000 2005 2010 2015 2020 AU 296 374 469 549 634 691 711 BE 50 117 274 422 560 683 711 DE 25 43 71 122 180 232 261 FI 93 139 206 210 229 243 246 FR 2606 3596 5010 8214 10961 13255 14092 GE 613 1187 2286 4014 5554 6809 7448 GR 674 1399 2909 5365 7270 9400 9737 IR 77 99 127 180 223 253 265 IT 8539 11242 16209 24337 27449 29801 30898 LU 3 5 11 18 23 27 29 NE 206 358 605 690 798 870 894 PO 436 669 1020 2049 3072 4040 4623 SP 7370 11744 19689 28331 33567 38707 39899 SW 227 309 391 378 403 422 426 UK 1443 1861 2359 3228 3832 4252 4417 Total 22660 33141 51636 78109 94755 109685 114656 BAU (MWh) 22660 33141 51636 78103 94727 109631 114579 Scenario 13 Obligation of moving from decentralised systems to centralised systems at a certain value of capacity (Portuguese regulation) With our figures this is not an efficiency option since the values we have favour decentralised AC. We cannot present figures displaying a positive impact of this measure.

37

Scenarii 14 &15 Reversibility is compulsory in all segments where reversibility is feasible Reversibility can be applied to many systems. There is no negative effect on Summer, the figures remain exactly the same. For winter, we make the results available for a team wanting to perform a comparison with other heating systems ,because the simple addition of kWh of any kind is a limited way of comparison. In our scenario 15, defrosting systems adopt the best technology available : defrosting on demand with heat exchanger wall temperature measurement. Reverse heating performances increase by 10%. Table 7.15 gives the commercial kWh gross figures.

Table 7.15 Total energy demand generated by AC in BAU, Scenarios 14 and 15

Electricity demand (TWh) Electricity and Gas

1990 1995 2000 2005 2010 2015 2020

Cooling function BAU (Electricity only)

22,879 33,683 51,636 78,103 94,727 109,631 114,579

Heating function BAU Without REV. 48,726 71,912

111,084 164,517 203,330 236,765 250,844

Heating function BAU With present REV. (Elec.) 7,274 11,390

18,894 28,913 35,875 42,333 45,040

Total BAU 78,879 116,985 181,614 271,533 333,932 388,729 410,463Non Rev Heating with compulsory reversibility SC14 48,726 71,912 111,084 159,012 165,882 157,637 119,919Reversible Heating with compulsory reversibility SC14 7,274 11,390 18,894 28,913 46,570 63,833 84,034Total Scenario 14 78,879 116,985 181,614 266,028 307,179 331,101 318,532Non Rev Heating with compulsory reversibility SC15 48,726 71,912 111,084 159,012 165,882 157,637 119,919Reversible Heating with compulsory reversibility SC15 7,274 11,390 18,894 28,688 45,169 60,932 79,381Total Scenario 15 78,879 116,985 181,614 265,803 305,778 328,2 313,879 General conclusions Fifteen technical scenarios have been simulated in every economic sector, country and for every type of systems up to year 2020. The very efficient scenarios are the following : 8 (combination of MEPS and building codes as in the USA), 3 (Eurovent part load MEPS), 2 (Eurovent full load MEPS). The maximum flexibility in demand is around 40TWh in 2020 if we start right now to apply the US Energy Efficiency standard, something which is far from the present prospects of regulation in Europe, except in two countries : Portugal and the UK. The Portuguese and British regulations would have a large impact if applied in the rest of countries. Significant gains can be obtained by introducing full and part load performance certification of equipment. A very strong commitment on full load only would achieve about the same gains at a very large cost.

Figure 7.3 Demand evolution, depending on scenarios

38

0

20000

40000

60000

80000

100000

120000

140000

1990 1995 2000 2005 2010 2015 2020

Sc.1Sc.2Sc.3Sc.4Sc.5Sc.6Sc.7Sc.8Sc.9Sc.10Sc.11Sc.12Sc.13Sc.14Sc.15

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8. EFFICIENCY RATING AT PART LOAD: AN IPLV FOR EUROPE

8.1 The importance and nature of part-load management measures

Importance of establishing a EU method about part load

Until now, as described in chapter 2.5, the chillers and other vapour compression cycles have been tested using the full load standard [CEN, 1997] for heating and [CEN, 1998] for cooling. This couple of standards has to be replaced soon, by its own revision [CEN, 2002]. Consumption is not governed by full load EER given in such a standard but by the average part load EER, called often a Seasonal EER (SEER). Such a figure is largely available on the US market and is called there an IPLV (Integrated Part Load Value). The one we are looking for should represent the universe of EU buildings and climates, hence the proposed name, the ESEER (European SEER). Only a ESEER largely agreed can be accepted as a basis for comparing chillers, or grading chillers from A to G in an undiscussable manner. We shall propose such an ESEER in the coming pages.

An extension to the new full load EU standard contains some information about part load testing, it

is an EnV, a provisional standard that could become a full standard (on which certification could be based) in case experience is gained about performing part load tests. In Italy, however, a part load performance standard has already been defined and accepted [UNI, 2002]. Part load performance is tested at different part load ratios, defined as the ratio of the cooling capacity of one stage to the full load capacity stage. The evolution of the efficiency with the part load ratio is still a subject for research and we have proposed here original results.

Given that the US-IPLV climatic conditions are not relevant for Europe, Italian manufacturers have

made a proposal for using the same methodology as for the IPLV but using different conditions for air and water condensing temperatures. The resulting index is called EMPE. The EMPE methodology is not different from the IPLV one. For a large set of modelled chillers, a comparison is drawn between EMPE and IPLV figures, leading to show the direct application of the IPLV to Europe would give overestimated values for the chillers’ seasonal efficiencies.

The goal of this chapter is to define an ESEER method that enables to calculate the seasonal

efficiency for all European chillers (centrifugal units are not treated explicitly in this document by lack of specific information but seem likely to be covered by the proposed method). The constraint is to minimize the testing time while ensuring maximum precision, it is to say that the error coming from the reduction of the data to single points should be inferior to the testing uncertainty. The new ESEER method is compared with the US-IPLV and EMPE proposal under both respects : time spent and accuracy.

The potential gain associated with part load management is high (for instance +30 % in EER, i.e. -

30 % in electricity consumption in our chapter 6.1 optimisation exercise). As long as a good method is not agreed, the gains and losses obtained by part load management can be mixed in some manufacturers documentation with more ‘artificial’ or ‘conventional’ gains and losses due to temperature conditions in testing. A good SEER definition is the essential tool for achieving actual and comparable gains not artefacts.

How to reduce the capacity of a chiller? The performance of each of the capacity steps will differ per se even if the operating conditions (entering air or water at the condenser and leaving water temperature) are identical. We need to describe the means of reducing the capacity to understand the reduced temperature and the part capacity behaviours. Both depend on the kind of compression circuit. The compressors treated hereafter are of three types, reciprocating, screw or scroll compressors. About part load behaviour of centrifugal chillers we give only qualitative indications.

Centrifugal compressors at part load

40

The efficiency of this kind of chillers at design point depends on : the size of the compressor (less stages = better efficiency because of intermediary losses). In a centrifugal compressor, before the impeller, the inlet vane guide enables to create more or less swirl to reduce capacity to match the load. This is a mechanical type of unloading. However, the two stages compressor enable to unload at a lower step. Speed of rotation : centrifugal chillers cannot be operated for small flow rates since the rotation speed needed would be too high. Part load : the surging phenomena occurs at low part loads, the flow comes back through the impeller leading to a cyclic phenomena badly known so that manufacturers forbid the chiller to work in these conditions. There are different modes for unloading the centrifugal chillers : prerotation inlet vane guides, variable speed. These two are the more common ones. However, when the load becomes inferior to the surging load, to enable cutting off the compressor, a hot gas bypass strategy is adopted leading to still poorer performances at very low loads. By associating in series two compressors the surge limit goes under the single compressor one (10% instead of 20 or 25% load). The part load performances seem always less than full load ones in what we have investigated.

Reciprocating compressors at part load The reciprocating compressor owns a spring valve at inlet and outlet. At suction, the inlet valve remains open as long as the pressure in the chamber is lower than the suction pressure. The valve at leaving opens only when the pressure in the chamber reaches the pressure of condensation. At that time, the end of the piston allows gas into the high pressure side. Thus, when external conditions vary, this compressor adapts its discharge and evaporative pressure to external conditions. This is the temperature aspect of part load.

The other treatment of part load is through capacity reduction. Mechanical realization of part capacity for a compressor with four pistons and two stages of compression is :

• only two pistons compress the refrigerant, the valves of both the others remaining open; the fluid which passes in the pistons in open position is pumped. This induces pumping losses.

• or only two pistons compress the refrigerant, the valves of both the others being closed; this induces a heating of the engine.

Screw compressors at part load In its basic configuration, contrary to the reciprocating compressor, the compressor has no means of adapting the pressure of exit of compression to the pressure of condensation. Thus, any difference between the discharge pressure and the condensation pressure is synonymous of energy losses.

For part capacity behavior, a slide valve or an equivalent steps system (corresponding to discrete steps for the bypass) is used to control the capacity. It makes it possible to shunt part of gases off the compression chamber to adapt the swept volume to the one needed to match the load. This slide valve (Figure 8.1) is generally controlled by discrete steps. In practice, the following stage is called variable Vi. The Vi is in fact directly related to the compression ratio since it is the ratio of inlet to outlet gas volumes of the compression chamber. Hence, the second valve enables to adapt the compression ratio to the condensing pressure for each capacity step while the first one enables to adapt the swept volume, and thus the cooling capacity of the chiller. The last option should be to use a variable speed drive for the motor. Some solutions with variable drive speed exist on the market. But the option is scarcely expanded because of its cost, given that reducing the speed of this compressor decreases the tightness of the lobes of the screw compressors, forcing the manufacturers to increase the full load speed to be able to reduce it at part capacity operation.

Figure 8.1 : slide valve position experiencing low Vi and high Vi (a), and Vi variable (b), from [PILL85].

41

Scroll compressors at part load The rationale is about the same one as for the basic screw : nor unloading is available for chiller applications, neither adaptation to varying condensing pressure. Generally, one uses several compressors in parallel on the same circuit and make them cycle. For two scroll compressors on the same circuit, two capacity steps plus the full load are available if nominal capacities of each of the two compressors differ.

Thus, different technologies are used to control part capacity stages. Depending on the kind of compressor circuit, one can find unloading by varying the number of available circuits or by varying the flow rate in one circuit. To perform this latter control of the refrigerant flow in the cycle, one can use variable speed drive (for screw chillers only in our scope), variable Vi unloading (for screw chillers only), unloading of multistage compressor (screw or reciprocating), or shutting down a compressor over two or more (sole option for scroll, available on screw and reciprocating as a supplementary mean).

Staging of Part capacity (control issues) Generally, capacity staged chillers are controlled using a water inlet or water outlet set point control with a dead band. Stages are successively triggered when water temperature increases and moves apart from the set point. The set point is always a control parameter to be entered by the user whereas dead-band can be fixed by the manufacturer or not. Figure 8.2 shows the control scheme for a five step capacity chiller on the inlet water temperature. Figure 8.2. Typical control of a 5 capacity stages chiller in the cooling mode. Part load (%)

Inlet water temperature °C

0 12 11 10 9 8 7 6

40

80

Dead-band 1°C

Dead-band 1°C

20

60

100

Given that chillers generally operate at full load and nominal inlet condensing temperature with a 5°C between inlet and outlet, the water temperature varies more or less between 6 and 8°C for all stages, depending on the water loop inertia and on the condensing temperature that modifies the cooling capacity of stages. We will assume here perfect control, the one represented by the scheme Figure 8.2, even if some experimental testing of dynamic capabilities of chillers have shown that chillers did not always behaved this way [AFCE, 2002]. However, dynamic testing installations are not available and would need long debates to be specified and then adapted by certifying laboratories. It has also been observed that most of the time, set point control temperatures were not respected, but differed by more or less 1°C and sometimes even more from user selected values. For correct measurements of inlet and outlet water temperature (like the ones used in standardised testing), either very long straight pipes are needed so that a homogeneous flow may be reached before measurements or, pieces of equipment have to be installed at the inlet and outlet of the evaporator to enhance the turbulence. On installed units such equipments are not installed, water temperature measurements do not correspond to real temperatures and the control behaviour can be far different from Figure 8.2. Cycling between stages at part load The following representation of the chiller performance when load is higher than the smallest capacity step is adopted : if the load lies between two capacity steps, the chiller will operate on each one of the two neighbouring steps ; the cooling load is the weighted average of the two steps cooling capacities CC for the same inlet condensing and outlet temperature. The corresponding operating times for each capacity step

42

enable to determine the electric power absorbed EP and thus the efficiency for each hour. Yearly efficiency is calculated with Equation (1).

∑

∑

=

== 8760

1

8760

1

ii

ii

EP

CCSEER

(1)

where CC is the cooling capacity and EP the electric power absorbed in each operating condition actually met.

When the cooling load is lower than the smallest capacity the equipment can deliver, the chiller operates only part of the time, thus fitting its cooling capacity to the load. In that case, each starting is an energy loss.

At each starting, the compressor has to establish the pressure difference between low and high pressure sides. The unit only begins to cool water when the average refrigerant evaporating temperature is lower than the average water temperature. Then, the superheating of the refrigerant has to stabilize : only at that time the full capacity of the step is reached. On the contrary, establishing the full electric power is quite instantaneous. This leads to an energy loss at the starting of the chiller. As a consequence of a review of all existing experimental evidence, we selected (figure 8.3) the Italian standard [UNI, 2002] Equation named (2) hereunder.

Figure 8.3. Ratio of part load efficiency to full load efficiency for the same inlet condensing temperature and outlet water temperature as a function of part load capacity in the same conditions.

The corresponding curve is represented on the figure 8.3, it corresponds to the following formula :

cyccycFL

FLFL C1.CCC

CCCC

CC

EEREER

−+= (2)

With Ccyc=0.9.

It is used here to compute the part load performances of single circuit units and the performances of multi-staged units when load is inferior to the smallest capacity step available.

Ccyc=0.9 is a proposed default coefficient. Supplementary work could be performed to set the experimental testing conditions enabling the calculation of this coefficient from the manufacturers specifications of the minimum water loop volumes and of the smallest capacity step available. High pressure control at part load The high pressure cannot go too low. It is generally maintained high enough by controlling the air flow rate for air cooled chillers. Classical control consists of maintaining the high pressure above 15 bar by cycling the

Degradation of the reduced efficiency versus the part load ratio (same sources temperatures )

00.10.20.30.40.50.60.70.80.9

1

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

Ratio of part load capacity to full load capacity (same source temperatures )

Rat

io o

f par

t loa

d ef

ficie

ncy

to

full

load

effi

cien

cy

43

fan or by switching one fan OFF (the first one being the better one since the whole area remains in use for the heat exchange) with a fixed dead band. Decreasing the flow rate at the condenser increases the high pressure and thus decreases the performance. The impact of this phenomena has been measured while carrying over a test campaign on a scroll unit. The unit is divided into 2 distinct and symmetrical refrigerating circuits. Each circuit has a tandem scroll compressor, which means two steps by circuit. Then, the capacity steps available are 100%, 75%, 50% and 25% ; in fact, due to the mechanical flow rate reduction, the 75%, 50% and 25% capacity steps, are slightly higher than this theoretical staging. Each circuit has 3 fans on its condenser (line configuration). Figure 8.4. Evolution of the chiller performances when reducing the condenser air flow rate presented under a reduced form (EER/EERnom, EP/Epnom, CC/Ccnom in terms of reduced flow rate)

The decrease of the efficiency while varying the flow rate is reported Figure 8.4. The chiller was operated at 50% load, one compressor in operation on each circuit. The 70% flow situation corresponds to 2 fans among 3 being ON for each circuit. Precise measurement of the flow rate was not available. The 50% flow rate corresponds to one fan functioning on one circuit and two on the second. The efficiency decreases with a square tendency when the air flow rate is reduced. The following point was to determine how the efficiency varies with the outside air temperature when the chiller operates at reduced flow rate : does the reduced efficiency increase with the same slope than with the full air flow rate ? Given the NUT-epsilon curve of the heat exchanger, at reduced flow rate, the increase of the efficiency with the outside air temperature decreases faster than at full flow rate. This fact is observed in reality, but in a more complex manner. At

full load, the high pressures are higher than at 50% load. Figure 8.5. Condensation pressure control effect on reduced EER (EER/EERnom) in terms of outside air temperature

Reduced performances while varying flow rate

0.7

0.75

0.8

0.85

0.9

0.95

1

1.05

1.1

1.15

1.2

0.5 0.6 0.7 0.8 0.9 1% of nominal flow rate

EER/EERnom

EP/EPnom

CC/CCnom

44

On figure 8.5 we can understand that the high pressure control will impact differently single circuit and double separated circuits units. For single circuit units, the 25% load would correspond to still higher ambient temperature for triggering fan cycling or reduction speed. Thus, the efficiency variation with outside air temperature is shown Figure 8.5 for 25% and for 75% load on a single circuit. On the contrary, for that double circuit unit, the 25% load point does not differ from the 50% since one compressor is in operation on one circuit, the 50% being strictly symmetrical. For double circuit and 75% load operation, one circuit is operated at full load while the other is operated at 50% load.

8.2 Is the IPLV approach directly applicable to European conditions? The percentage of operating hours spent at each part load condition, given in our description of US-IPLV, chapter 2.5, (1% etc…) is intended to be representative of the US climate and buildings but not of the European ones. Further to this, an analysis of the method shows that the ARI part-load temperature testing points are "sized" to be "representative" of US buildings (cooling even in negative Celsius temperatures, for instance- as can be shown by drawing the loads in terms of outside temperatures). Thus not only has the load been varied as in the draft CEN part-load test standard which is currently under discussion but also the temperatures.

Buildings used in deriving the US-IPLV This standard covers all the tertiary sector buildings for air conditioning application on the whole US territory. A single building has been “averaged” to be representative of buildings of 29 cities5 chosen to be representative of the places where chillers are installed in the US.

Four building groups have been identified depending on the occupation scenario and the possibility to use free-cooling or not :

• Group 1, occupation 24h/day : 7days/week, cooling above –17.2 °C,

• Group 2, occupation 24h/day : 7days/week, cooling above 12.8 °C, free-cooling between –17,2 and 12,8°C.

• Group 3, occupation 12h/day : 5days/week, cooling above –17.2 °C,

• Group 4, occupation 12h/day : 5days/week, cooling above 12.8 °C, free-cooling between –17,2 and 12,8 °C.

Climate used in IPLV derivation 5 [ARI98] states that these cities represent 80% of the installed chillers in the US.

Reduced EER versus OAT for different load ratios

1,00

1,10

1,20

1,30

1,40

1,50

1,60

1,70

15 17 19 21 23 25 27 29 31 33 35

OAT (°C)

DOE2 curve

FL curve

PL curve, singlecircuit 50 %

PL curve, singlecircuit 75%

PL curve, singlecircuit 25 %

PL curve, singlecircuit VSDF

Without highpressure control

45

The mean climate used to perform the bin method is an average of the 29 cities climates. The climatologic data were averaged without any weighting of energy or capacity installed in each city. Coming from this average climate, an occurrence curve of dry bulb and wet bulb air temperature is drawn by 5 °F (2.8 °C) bins, between –17,2 °C and 35 °C. The water temperature is deduced from the wet bulb temperature using an added 8 °F (4.4 °C) approach.

Building cooling load calculation in US-IPLV The details of the calculation are not explained. The only indications are given in the text. Internal loads represents 38% of the total load above 12.8°C. The building load experiments either a 20% of maximum load at 0°F (–17.2°C) for groups 1 and 3 either is null under 12,8°C for groups 2 and 4. The nominal full load sizing for the four groups correspond to the highest bin, temperature higher than 95°F, or 35°C. The load curve of group 1 is given figure 8.6. Each group is weighted in function of its relative weight in the US from a statistical study.

Calculating US weighing coefficients The following calculations depend now on the group chosen. Then, the results are weighed by the representative coefficients of the groups amongst US building studied.

• Taking group 1 as an example, the load curve is multiplied bin by bin by the number of hours experienced in each bin considering the average climate.

Then, one obtains the energy needs curve (figure 8.6). The ARI 550/590 unit for energy is the ton-hours. What is shown is actually the product of the hours by bin and of the normalized capacity in %. The real unity is the hour but this name enables to remember that it characterizes the repartition of the energy to be delivered versus outdoor temperature.

Figure 8.6 : annual cooling needs as a function of outside air temperature, group 1, from [ARI98].

• On this curve, four integration intervals are then defined:

The sizing interval (for temperature higher than 95 °F : 35 °C) and the three other intervals : [0,55], [55,75], [75,95]. For each interval, one integrates the energy curve that will give the energy weight of the four testing points called respectively A, B, C and D values.

Interpolation scheme needed to reduce testing time If the unit cannot unload to one of the specificied steps, two possibilities exist : either it can unload at a lower capacity than the missing step or it cannot. If it can, the missing efficiency will be obtained by interpolation of the two closest embedding efficiencies available within testing point according to ARI testing load temperature curve.

46

Figure 8.7 illustrates the interpolation procedure for unit N°7. The graph differs from the representation used in the IPLV standard for part load. The load ratio in % is related to full load and 35°C inlet air temperature. The solid and very black line is the ARI assumed load/temperature curve. The less solid black line is determined by the capability of the four steps of the chiller to reach the four specified temperatures. Three straight lines representing the capacity stages relative to full load 35°C capacities are drawn for the stages 1, 2 and 3).

Figure 8.7 : Interpolation procedure illustration

One example of the interpolation can be described with the help of figure 8.7 : when the outside air temperature is 18°C and the load between about 30% and 62%, the chiller will cycle between the 50% and 25% stages. Thus the efficiency of the 50% point 18°C should be calculated as the weighted average of the two points: [18, first stage] and [18, second stage]. The corresponding point is located at the crossing point of the ARI curve and the horizontal plain arrow at 18°C and 50% load ratio. To avoid the multiplication of the number of testing points, the ARI procedure uses only the testing points (full circles) to perform interpolation of the efficiencies for specified load points. Thus instead of weighting the two previously mentioned points, the standard proposes to weight the [18, first stage] and the [26, second stage] points. Thus, the 50% point efficiency is underestimated in that case since efficiency of the [26, second stage] point is lower than the [18, first stage] due to temperature decrease. The consequence is that for that unit number 7, the global seasonal figure is underestimated by the interpolation procedure whereas, for a continuous control screw unit, the IPLV exact figures would be obtained, the capacity step chillers being penalized for not supplying continuous unloading. The method is interesting since it enables to reduce the testing points number and its effect will be discussed further hereunder. EMPE: an answer to a need for a European weighting with IPLV-like testing The first remark that led to the EMPE [AICARR, 2001] Italian proposal is that the operating conditions are rather different from Southern Europe conditions. And even, if Northern Europe countries may need air conditioning in summer, it cannot be said that Italy would need air conditioning at 12.8 °C as normal operating conditions.

10

15

20

25

30

35

0% 10% 20% 30% 40% 50% 60% 70% 80% 90% 100% Part load ratio (%)

Outside air temperature (°C)

ARI aircondensationcurve75

50

25

47

Therefore, AICARR proposed a new energy index, named EMPE (Average Weighed Efficiency in Summer regime in Italian) directly deriving from IPLV, with different energy weights and, in particular, with different temperatures at the condenser inlet, fitter for the European climate and requirements in the air conditioning field.

The EMPE formula is absolutely similar to IPLV, but the values of energy weights and inlet temperature to the evaporator and the condenser are those indicated in table 8.1:

Table 8.1: calculation conditions for EMPE

The AICARR proposal, EMPE was not based on a sufficiently large climatic and technical investigation. Its strength (being very close to the existing US method, which aggregated many factors) was also its weakness. We had the opportunity to go further by constructing a data base of EU chillers at part load, understanding better part load, and proposing two separate methods, one for part load reporting and certification, the other one for the computation of SEER.

Reduction of EMPE or IPLV to 2 points with extrapolation The following method has been proposed by one manufacturer. It is an amendment to IPLV or EMPE. Two testing points only are performed, the first one at 100% load and design temperature and the second one at 50% of the load and the associated reduced temperature. The EER values at 75% and at 25% of the load, necessary to define the index, would be obtained by interpolations and extrapolations on the IPLV or EMPE curve. Substantially it is assumed that the efficiency changes linearly with the load (the temperature also decreases as in the IPLV or EMPE or at other conditions). The manufacturer could then decide if making 4 tests or only 2.

The EER value at 75% of the load is calculated as follows:

48

2%50%100

%75EEREEREER +

=

The EER value at 25% of the load can be calculated as follows:

)( %75%50%50%25 EEREEREEREER −+=

This system is representative when a control step is placed at 50%. We have tested this simplification extensively on a set of chillers (table 8.2).

Table 8.2 : Comparison of the proposed 2 points methodology for scroll air condensing units with the EMPE

Air cooled scroll chiller n° 5 6 11 13 14 15 2 3 EMPE 3,35 3,22 3,50 3,59 3,32 3,39 2,80 3,39

2 points method 3,82 3,42 3,92 3,85 3,16 2,91 3,26 3,48 Relative difference with EMPE 14% 6% 12% 7% -5% -14% 16% 3%

The uncertainties that have been generated are too large compared with the accuracy expected for the seasonal index. Moreover a bias is introduced in the classification : some always loose, some always benefit. This methodology suffers the same bias that was introduced by the interpolation process, again increased. The evolution of part load efficiency at reduced temperature cannot be modelled simply by a linear regression : it depends on the unit, even for the very commune air scroll range. Moreover, practical limits as well as the impossibility to predict cycling keep us from recommending that method.

8.3. Construction of a data base of EU chillers at part load –understanding part load

Testing conditions and available testing results Original knowledge has been generated during the “Joint project” of EDF R&D facility and manufacturers from Eurovent wanting to promote part load performance. The main tool used was actual testing of EU equipment but a number of group meetings allowed to build a common thinking frame. The technical description of the chillers tested follows on tables 8.3 and 8.4, split by condensation type.

Table 8.3. Tested air-cooled chillers Name Type Circuits Compressors Available Stages N° 5 Scroll 1 2 3 N° 7 Scroll 2 4 4 N° 8 Herm rec 2 2 2 N° 9 Scroll 2 4 4 N° 2 Screw 2 2 Partially continuous

Table 8.4. Tested water-cooled chillers

Name Type Circuits Compressor Available Stages N° 1 Screw 2 3 8 N° 3 Screw 2 2 4 N° 4 Scroll 2 4 4 N° 6 Screw 1 1 Continuous

For all the tested chillers, some common testing points were made according to either the ARI or the

EMPE conditions depending on the manufacturer will. For all chillers, a supplementary point was added to fulfil the CEN EnV requirement : nominal inlet condensing temperature (35°C for air and 30°C for water) and 50% load ratio referred at this nominal inlet condensing temperature [CEN, 2002]. For chillers n° 2, 3, 4

49

and 8, only IPLV or EMPE points plus the CEN one were available. For the others as many testing points as desirable have been obtained. In all circumstances a simple model has been used to draw the performance maps from existing testing points.

Impact of load reduction on the efficiency – a reporting format proposed to Eurovent

The main finding is that a percentage (like 50%) is not enough to characterise a part load behaviour of a chiller. It is so when there is one single compressor per chiller, or various identical circuits. A significant market share of chillers have various compressors and a complex circuiting, leading to improved part load performance. But a given part load regime has to be defined by the actual status of each piece of equipment.

For discrete stages chillers, it would be easier to describe performance at a given stage not at a given

percentage. For the very few continuously controlled chillers, fours stages can be defined in terms of input. Since temperature and load can be tested independently and recombined, there is no need for combined testing (like IPLV).

About certifying Part Load : what the manufacturers give to their customers is a « map » of

performance, not only values at the four arbitrary percentages and temperatures, plus the final Eurovent grading when it is available, based on a SEER. The customer can rely on the Eurovent SEER computed from this map … or compute its specific SEER for its specific demand. No need to test every condition reported in the “map”: the benefit of Eurovent is the fair and independent choice of a few points on the map, as usual, and the associated independent testing.

We arrived also at applicable conclusions on the way to report the SEER in the Eurovent directory.

We started from HSEER, the DOE reference that we generated. It is proven that each set of outside conditions (for each sector, climate, type of chiller, type of secondary system) can be reduced to four or five external conditions without loss of accuracy. The ESEER index proposed here is a set of 4 conditions given for E.U. as a whole, but there can be as many similar indices as specific demands: sector, country, etc.

We have introduced a format for the description of the stages of a chiller, like in table 8.5 and following, suitable for

Eurovent specification. For each stage, the manufacturer has only to declare which piece of its equipment is operating and to indicate CC , the cooling capacity and EP, the electric power absorbed. The certifying body has only to check a few of the values, selected in the same conditions as usual. Note that this procedure is in fact already used for some chillers with various speeds, namely “low noise” chillers with the possibility of reduced fan speed.

Table 8.5 : Part load performance of water cooled scroll chiller N°4, as could be reported in Eurovent part load

certification scheme N° 4 // WT : 30°C STAGES 1 2 3 4

Compressor 1 0 0 0 1 Circuit 1

Compressor 2 0 1 1 1 Compressor 3 0 0 1 1

Circuit 2 Compressor 4 1 1 1 1

EP (kW) 8,80 17,60 27,17 38,27 CC (kW) 37,50 78,00 112,50 150,00

EER 4,27 4,47 4,12 3,92

Now we shall present examples of the proposed procedure. Let’s note that it is far easier to analyse the part load behaviour of water cooled than it is for air cooled chillers. Indeed, the air cooled chiller stage efficiencies can suffer different fan pattern and/or circuit separation that do not infer for water cooled chillers. The chiller part load behaviour is described first for water cooled units, then air cooled units. Water cooled chillers –experimental results

50

The overall performance improvement (or degradation) at part load (temperature effects being substracted) is given on figure 8.8 for the four tested units.

Figure 8.8. Reduced efficiency while decreasing part load ratio (same source temperatures) for the testedwater cooled chillers

Let’s give more explanation about two of the tested chillers, N°4 and N°1, as examples of real life issues. Chiller N°4 is a two circuit four scroll compressor chiller, with the same symmetrical tandem on each circuit. The efficiency increase show that at 50%, one compressor by circuit is activated. At 75%, one circuit is at full load and the other at half load. At 25%, only one circuit works at half load. Logically, at 25% and at 50%, the symmetry of the chiller would impose identical performances. The bias can come from many causes, one being the specific configuration of the plate heat exchanger : the two distinct refrigerant circuits use the same brazed heat exchanger in order to cut the costs. This complex part load behaviour can be summed up under one form per temperature – see Table 8.6, or one single table with all temperatures –table 8.23.

Table 8.6 : Part load performance of water cooled scroll chiller N°4 N° 4 // WT : 30°C STAGES 1 2 3 4

Compressor 1 0 0 0 1 Circuit 1 Compressor 2 0 1 1 1 Compressor 3 0 0 1 1

Circuit 2 Compressor 4 1 1 1 1

EP/EPFL 23% 46% 71% 100%CC/CCFL 25% 52% 75% 100%

EER/EERFL 109% 114% 105% 100% Let’s note here that the presentation under this format ensures that the manufacturer has consciously chosen this staging as optimum and hopefully that it has been prioritised as factory default control parameters. Now let’s consider chiller N°1 (Table 8.7) : it is a double circuit water screw chiller with one compressor by circuit. One can see easily the difference between the two type of unloading, symmetrical, for higher than 50 part load ratios and on a single circuit, reducing the refrigerant flow rate at the minimum for part load ratio smaller than 50%. In that case, the efficiency decreases somehow faster. Stages configurations and performances are gathered Table 8.5 under the form proposed to Eurovent.

Table 8.7 : Part load performance of water cooled scroll chiller N°1 N°1 // WT : 30°C STAGES 1 2 3 4 5 6

Circuit 1 Compressor 1 0% 0% 0% 59% 101% 100%

Reduced efficiency of the part load stages for water cooled chillers

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

0 0.2 0.4 0.6 0.8 1 1.2Part load ratio

N° 4N° 1N° 3N° 6

51

Circuit 2 Compressor 2 51% 58% 100% 56% 56% 100% EP/EPFL 33% 36% 51% 70% 87% 100% CC/CCFL 25% 28% 49% 58% 79% 100%

EER/EERFL 74% 78% 96% 82% 91% 100% It clearly appears that 3 stages are used for compressor 2 (100, 56 and 50) and 2 only for compressor 1 (100 and about 60). The percentage for each compressor corresponds to the ratio of the cooling capacity of the compressor to its full load capacity (half the chiller capacity). Air cooled chillers –experimental results The overall performance improvement (or degradation) at part load (temperature effects being substracted) is given on figure 8.9 for the five tested units. Figure 8.9. Reduced efficiency while decreasing part load ratio (same source temperatures) for the water cooled chillers

Let’s give more explanation about two of the tested chillers, N°5 and N°2, as examples of real life issues. Table 8.8 gives testing results on chiller N°5 : it is a single circuit unit with an asymmetrical scroll compressor tandem, which means three capacity steps. The increased efficiency with reducing the refrigerant flow rate from stage 2 to stage 1 is counterbalanced by the relative increasing weight of the fan consumption.

Table8.8: Part load performance of air cooled scroll chiller N°5 N°3 // OAT : 35°C STAGES 1 2 3

Compressor 1 1 0 1 Compressor 2 0 1 1

Circuit 1 Fan 1 1 1

EP/EPFL 38% 52% 100%CC/CCFL 46% 64% 100%

EER/EERFL 120% 124% 100% Now let’s consider N°2 in table 8.9 : it is a screw double circuit unit. Refrigerant circuits are separated. There is one screw compressor by circuit. Each compressor can unload partly continuously from about 100% to 75% and then two supplementary stages at 66% and 33% are available for each one. Only the four tested stages are reported. The 50% and 75% capacity stages cannot be allocated to each circuit. Only the 25% can. This part load behaviour just confirms that unloading on a single circuit with a slide valve is very inefficient as compared to the full load efficiency of the screw.

Reduced efficiency of the part load stages for air cooled chillers

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

0 0.2 0.4 0.6 0.8 1 1.2

Part load ratio

N° 5N° 7N° 8N° 9N° 2

52

Table 8.9 : Part load performance of air cooled screw chiller N°2 N°2 // OAT : 35°C STAGES 1 2 3 4

Circuit 1 Compressor 1 30% ? ? 100% Fans 3 3 3 3

Circuit 2 Compressor 2 0% ? ? 100% Fans ? 3 3 3

EP/EPFL 28% 55% 64% 100%CC/CCFL 15% 50% 71% 100%

EER/EERFL 52% 89% 111% 100%

8.4 Derivation of a new SEER method (ESEER) Given the complexity of the subject, the EECCAC group adopted the building simulation tool DOE2 to simulate representative buildings of the European stock market. Some studies are available in Europe giving the description of the commercial building stock and a very few countries have also developed the buildings within simulation tools. However, when multiplying the simulation cases including building types (offices, malls, hostels, hospitals, administration …), the climatic conditions and the different systems, it led to an incommensurable number of simulation, without mentioning the number of buildings to be entered in the used building code. We had to make some decisions.

The simulations leading to the reference values of SEER (HSEER) Two buildings were simulated on computer, but buildings that do exist : an office and a commercial mall. For each one, three climates have been simulated, adopting different envelope characteristics when moving the building around Europe. The different systems identified in the stock and market study have been simulated. CAC air and water distribution equipments have been simulated using the European average efficiency values.

Hour after hour, the simulation uses then the characteristics of the real chillers modelled to compute the exact yearly performance index : the HSEER (Hourly SEER), used then as a reference for other methods. At each hour the outside enable to calculate all known stage capacities and respective electric powers, including the high pressure control impact on each stage. Then the load is compared to each stage capacity. If the load is lower than the smallest available capacity step, the cycling formula enables to calculate the electric power. Otherwise, the weighting of electric power of each stage is found by the expression of the weighted average.

In fact this computation has been performed in two steps : first, chiller hourly load curves were extracted as well as the coincident hourly OAT and the specific humidity. Then those conditions are used a number of times with different chillers’ and equipment quality. The different climatic conditions and central air conditioning systems available are described Table 8.10.

Table 8.10. Available hourly load curves

Sizing issues for chillers rating as shown by the simulation of the buildings The Milan CAV hourly load curve is presented Figure 8.10. The tendency that links load and temperature is very clear even if an important scatter is observed.

Figure 16. Milan office building CAV system hourly load curve

Office building Climate

System type London Milan Seville CAV

CAV-FC VAV

FCU 4P

53

When trying to calculate air cooled chiller performances from hourly load curve, two problems appear that will show up two significant limitations of the ARI method.

The load curves must enable to calculate the consumptions for all chillers. Thus, the load must be divided by the sizing load, so that all the chillers may be compared on the same load and temperature repartition. In the ARI standard, that point is solved by assumption since all the straight line load curves have the same maximal temperature of 35°C that also corresponds to the maximum load.

It clearly appears Figure 8.10 that the ARI hypothesis is not verified. The explanation is that even if temperature is correlated to the load, other load pattern intervene as the solar loads, the thermal inertia and the dehumidification loads that are the sources of the non explained variance by the load and temperature correlation. This is the first limitation of ARI sizing assumptions.

So the real optimal design rule is : the maximal chiller capacity and the corresponding temperature corresponds to the (load, OAT) couple that enables the chiller to cover all the cooling needs. The capacity variation with OAT of a perfectly sized (500 kW, 30°C) chiller with this simple law has been drawn Figure 8.10 and shows sizing is correct. Of course, if maximal load were at lower temperature, it could happen that the sizing could lead to non-satisfied needs ; an iteration process on hourly load and chiller capacity has been adapted to make sure that the cooling capacity is enough all the year long.

The sizing realized for the three climates led to 30% constant oversizing for the 3 climates for the office building. For real world installations, security coefficients are generally applied to the simplified sizing method leading to huge oversizing up to 100%. Given the part load characteristics of the chillers, it seems obvious that consequences for the seasonal efficiency will also be very important.

The seasonal efficiencies for the MILAN CAV load curve are presented Table 8.11 for 30% and 60% oversizing for the air cooled chillers N°7 and N°2 .

Table 8.11. Impact of oversizing on seasonal performances for Milan CAV hourly load curve SEER values

Oversizing N°7 N°2 0% 3.81 3.12

30% 3.83 2.81 60% 3.76 2.60

The differences are limited for the air scroll chiller N°7, because the reduced part load efficiency is still higher than 1 at 25% and hardly lower than 1 for the 50% load reduced efficiency. On the contrary, for the screw chiller N°2, a sharp efficiency degradation with the load had been noted. These results confirm that the

Hourly load curve

050

100150200250300350400450500550

10 15 20 25 30 35OAT (°C)

kWh CAV-MILAN

54

sizing is a key factor for seasonal performances analysis. And it also shows that no optimum sizing can be done without the exact knowledge of the chiller part load performances.

In the ARI methodology, for reducing the load curve and temperature occurrences to 4 points, only the 100% and more than 35°C OAT couples are kept. Whereas in our calculation methodology, all the points between the full load stage and the step immediately inferior are shared between the two steps, giving weight to the full load. The ARI methodology of reduction would lead almost to a null energy share for the full load stage, which is not true for staged chillers but approaches the truth for continuous control chillers. This is the second limitation of ARI sizing assumptions.

Reduction of European hourly load curves to a set of four conditions (based on the example of Milano) To reduce the load curve to 4 points, it is supposed that a virtual chiller, with 4 capacity steps at 25%, 50%, 75% and 100% of the design load is operating. The Milan-CAV load curve is used to illustrate the methodology. But this methodology can –and will- be applied to all conditions obtained by simulation, leading to the possibility of a four points representation of any condition or of the EU average of operating conditions.

STEP 1 of the load curves reduction process The sizing described above enables to transform Figure 8.10 with % load instead of kWh on the Y axis. Then, the load curve is put under a grid format. On the X axis, the outside air temperature is binned. On the Y axis, the load ratio is binned. In each rectangle, the relative kWh are added, giving a repartition of the cooling energy needs on the grid (Figure 8.11).

Figure 8.11. Grid representation of the Milan-CAV load curve (load ratio bin length : 0.05, temperature bin length : 2°C)

The average temperature is kept for each column and called hereafter binT(k). For each line, the average load ratio is kept and called here after binL(i), where k varies between 1 and K, the number of temperature bins and i varies between 1 and I, the number of % load bins.

STEP 2 of the load curves reduction process A statistical reduction method enabled us to get, from the hourly load curve, the representation of a discrete load curve and a discrete weighting curves.

Figure 8.12 gives for instance the discrete load curve for the hourly simulation of the office building in Milan. It shows that it is not far from a straight line, as assumed in the ARI standard. However there would be no weight for the 100% stage applying the ARI methodology from now on, as already explained.

Figure 8.12. Reduced load curve for the office building, Milan-CAV

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Figure 8.13 gives the representation of the energy weight of each class as a function of temperature classes. It is equivalent to the ton hour curve given in the ARI standard for Group 1 (Figure 8.7). It can be seen that even for Milan, the chiller load and the associated weighting energy coefficients are still not null at temperatures as low as 15°C.

Figure 8.13. Reduced weighting curve for the office building, Milan-CAV

STEP 3 of the load curves reduction process The ARI methodology load curve reduction would not give any weight for the full load point. We know this is not true and we have used a simple 4 stages chiller to determine the weight of each stage and then the corresponding operating temperatures.

At this level, for each capacity step, we have obtained weighting coefficients that represent the energy weights to be associated to the % load required (25, 50, 75 and 100) at each temperature bin. Each stage weight and OAT repartition is represented Figure 8.14.

Figure 8.14. Reduced weighting by stage under the PT format, method 3rd Step, for the Milan-CAV office building hourly load curve

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So now, the weighting coefficients are known :

SEERP= [0.2024 0.4272 0.3369 0.0335]

STEP 4 of the load curves reduction process The average temperature for each stage is calculated and the following operating temperatures are found :

SEERT=[17.9997 23.0611 28.0627 31.0514]

The results for the reduction for the Milan office building load curve are presented Table 8.12.

Table 8.12. Reduction of the chiller hourly load curve for the office building in Milan using the CAV system for air cooled chillers

Part load (%) Reference (nominal full load) Inlet air temperature (°C)

Energy weighting coefficients A, B, C, D

100 31.2 3%

75 28.0 34%

50 23.1 43%

25 18.1 20%

Results for more extreme weather conditions (London, Seville, different distribution systems) The hourly calculation methodology has been applied to two different load curves among the twelve available (Table 8.10) :

• CAV-FC system in Seville • CAV system in London

However, in order to separate the quality of the reduction by itself from the non linear models representing the chillers, the seasonal performances are calculated successively for the 4 following configurations :

• Without cycling, without high pressure condensation control • Without high pressure condensation control, with cycling • Without cycling, with high pressure condensation control • With both phenomena

The reduction results of the two specified load curves are presented Table 8.13. The two extreme load curves reduction presented show that both weighting coefficients and temperature conditions greatly vary with climatic conditions and systems.

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Table 8.13. Reduction of the chiller hourly load curve for the office building in Seville using the CAV-FC system for air cooled chillers and in London, using the CAV system

Seville CAV-FC London CAV

Part load (%) Reference (nominal full load) Inlet air temperature (°C)

Energy weighting coefficients A, B, C, D Inlet air temperature (°C)

Energy weighting coefficients A, B, C, D

100 36.7 4 % 27.6 1 % 75 32.1 48 % 24.8 10 % 50 27.4 37 % 20.9 42 % 25 22.8 11 % 17.1 47 %

The results of the reduction methodology are presented Table 8.14. The nominal full load efficiency (at 35°C) is reported for each chiller and so is the hourly seasonal efficiency ratio (noted HSEER for hourly), the reduced index figure (noted ESEER) and the relative efficiency difference between the two seasonal figures. We translate the information in terms of ranking : chiller ranked 1 is better than chiller ranked 2, and so on.

Table 8.14. Accuracy of the reduction for the tested air cooled chillers Conditions Seville CAV-FC load curve London CAV load curve

Chillers N° 5 N° 7 N° 8 N° 9 N° 2 N° 5 N° 7 N° 8 N° 9 N° 2

EER 2.18 2.59 2.51 2.47 2.93 2.18 2.59 2.51 2.47 2.93 EER ranking 5 2 3 4 1 5 2 3 4 1

HSEER 3.01 3.38 2.73 3.09 3.13 4.08 4.59 3.30 4.21 3.19 ESEER 3.05 3.44 2.73 3.14 3.15 4.08 4.60 3.30 4.22 3.31 No cycling, no fan cycling Relative deviation 1.2% 1.8% 0.0% 1.6% 0.7% 0.0% 0.2% 0.0% 0.2% 3.9%

HSEER 2.93 3.34 2.69 3.06 3.11 3.70 4.43 3.08 4.11 3.14 ESEER 3.00 3.43 2.72 3.14 3.15 3.88 4.58 3.20 4.22 3.31 Cycling only Relative deviation 2.4% 2.6% 1.1% 2.4% 1.3% 4.7% 3.3% 3.7% 2.7% 5.5%

HSEER 2.99 3.35 2.71 3.05 3.00 3.80 4.26 3.05 3.91 2.96 ESEER 3.03 3.40 2.73 3.11 3.10 3.72 4.31 3.01 3.90 3.07 Fan cycling only Relative deviation 1.4% 1.4% 0.8% 1.9% 3.4% -1.9% 1.4% -1.4% -0.4% 3.5%

HSEER 2.90 3.31 2.66 3.03 2.98 3.46 4.12 2.86 3.82 2.92 ESEER 2.98 3.39 2.71 3.11 3.10 3.56 4.31 2.93 3.90 3.07 Cycling and fan cycling Relative deviation 2.7% 2.3% 2.0% 2.6% 4.0% 2.9% 4.5% 2.4% 2.0% 5.1%

HSEER ranking 4 1 5 2 3 3 1 4 2 5 ESEER ranking 4 1 5 2 3 3 1 4 2 5

The following remarks comment the key points of Table 8.14 :

• The full load nominal efficiency appears to be a poor indicator of the seasonal efficiency. No policy measure should be based on it only as far as capacity staged or continuous capacity control chillers are concerned.

• For both conditions, the ESEER follows the HSEER classification of seasonal efficiencies. However, the ESEER values are always higher than the HSEER values by 2% to 5.1%.

• For the two different load curves, it appears that the chiller classification is largely modified ; it shows the important impact of load on the seasonal efficiency. The chiller N°2 is the first for the EER sequence, the third for the Seville sequence and the last for the London sequence.

• The relative maximal bias created for Seville by the methodology reduction is [+ 2%, + 4%]. The relative maximal bias created for Seville by the methodology reduction is [+ 2%, + 5.1%].

• Cycling and fan cycling impacts cannot be neglected for the London curve conditions as it is shown

by the HSEER evolution. However, both impacts are logically very low for the CAV-FC load curve in Seville, conformingly to table 9 reduction results.

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Table 8.15 below enables to check that the load curves selected for the sensitivity analysis represent indeed extreme conditions for temperature and load. For the London CAV system, the temperatures are the lower (except for the FC4P system) and weighting coefficients maximum at low loads. For the Seville CAV-FC system, respective reverse conclusions on temperatures and load can be made.

Table 8.15. Applying the reduction methodology to the set of available load curves Temperatures (°C) Weighting coefficients

Load 100% 75% 50% 25% 100% 75% 50% 25% Climate System London CAV 27.6 24.8 20.9 17.1 0.7% 9.7% 42.5% 47.1%

CAV-FC 27.6 26.1 24.3 22.4 4.2% 26.7% 37.6% 31.5% FC4P 27.6 24.6 20.1 16.1 0.5% 8.7% 48.5% 42.3% VAV 27.6 25.6 22.4 17.6 1.1% 7.7% 29.1% 62.1%

Milan CAV 31.2 28.0 23.1 18.1 3.6% 33.9% 41.7% 20.8% CAV-FC 31.2 28.0 24.8 22.0 5.7% 54.4% 31.1% 8.8% FC4P 31.4 28.1 23.1 17.7 3.1% 32.0% 40.5% 24.3% VAV 31.6 28.9 24.5 19.1 2.6% 30.7% 39.5% 27.2%

Seville CAV 36.7 32.1 26.3 19.8 3.5% 38.2% 39.1% 19.2% CAV-FC 36.7 32.1 27.4 22.8 4.4% 47.5% 37.3% 10.7% FC4P 36.9 32.3 26.5 19.2 2.8% 35.3% 40.2% 21.7% VAV 37.2 33.4 28.0 21.1 1.6% 30.7% 43.9% 23.8%

Thus it can be concluded that the methodology proposed is a qualified tool to classify the air cooled chillers at the condition to respect seasonal efficiency classes wide at least of 5% of the market average ESEER absolute figure. Extrapolating to the European stock of chillers in use Results of this weighting (described in chapter 5) are given Table 8.16 for air cooled chillers. Table 8.16. Cooling energy needs national weighting coefficients for air cooled chillers for CAV and FCU (WC :

weighting coefficient) Country Aus Bel Den Fin Fra Ger Gre Ire Ita Lux Neth Por Spa Swe UK

CAV 0.8% 0.3% 0.1% 0.2% 9.9% 3.2% 5.4% 0.1% 38.0% 0.0% 0.7% 1.1% 37.3% 0.4% 2.5%WC FCU 0.8% 0.4% 0.1% 0.3% 9.0% 3.2% 4.9% 0.1% 34.6% 0.0% 0.8% 1.0% 41.4% 0.6% 2.8%

Southern Europe country visibly represent most of the cooling energy needs in Europe. We can reduce this information to a set of 6 coefficients that will be used to weight the SEER obtained from London, Milan and Seville load curves. Final weighting coefficients for the available load curves types are presented Table 8.17.

Table 8.17. Cooling energy needs hourly load curves weighting coefficients for the air cooled chillers, for CAV and FCU, free cooling and VAV specific applications (WC : weighting coefficient)

Climate London Milan Seville CAV 5.3% 16.1% 18.4% FCU 8.2% 22.2% 29.8%

Free Cool. 6.5% 41.6% 51.9% WC

VAV 10.6% 40.3% 49.1% CAV (air distribution) and FCU (water distribution) systems enough to represent the present stock. Figures of the two first lines of Table 8.17 for CAV and FCU are to be understood as a complete set of coefficients for 6 load curves. The sum of the 6 weighting coefficients equals 1. For the values for CAV+Free Cooling and VAV applications, only the air distribution systems are concerned . The 3 weighting coefficients issued from CAV market shares are here to generate scenarios of improved efficiency.

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At last, these coefficients enable to weight the results of Table 8.15 and to present final ESEER average European conditions in Table 8.18. The values of the two first arrays are our proposal for EU chillers rating. The other results show how free cooling and VAV application could change those recommended values.

Table 8.18. Application of the method to derive a ESEER for air cooled chillers used with the free cooling or VAV options

ESEER Free-cooling VAV

Part load ratio Temperatures Weighting coefficients Temperatures Weighting

coefficients Temperatures Weighting coefficients

100 34.0 3% 33.9 4% 34.6 2% 75 30.1 33% 30.1 45% 31.2 30% 50 24.7 41% 25.5 35% 26.3 41% 25 18.6 23% 21.0 15% 20.1 28%

The differences for the VAV application for weighting coefficients can be neglected while each stage temperature increases by about 1°C. For free cooling, the lower stage temperature increases while weightings move towards higher loads. For both options, the average coefficients only slightly moves, confirming that the system driving efficiency factor is respectively the load avoided for the free cooling option and the fan consumption avoided for the VAV option. In a similar way the values have been defined and validated for water cooled chillers. However, for water cooled units, the water temperature at the condenser inlet depends not only on the OAT but also on :

o the condensing water flow rate,

o the tower performance curves,

o the specific humidity.

Only the open type towers have been considered here, since they represent 80% of the European stock, despite of the Legionella disease that certainly greatly modified the sales. Within the ARI standard, it was supposed that the approach (in that case defined as the temperature difference between the inlet water temperature at the chiller and the air wet bulb temperature) was constant for all conditions of operation, which is false. Here, real towers have been sized using the cooling towers electronic catalogue of the leading manufacturer, completed with the NUT-epsilon heat exchanger theory, considerations on control of equipment and recent correlations. On Figure 8.14, for a specific screw chiller extracted from the manufacturer catalogue, the part load ratios versus the condenser inlet water hourly temperatures are drawn. The load curve is the Milan CAV one. The effects of the chosen control scheme clearly appear : the inlet condenser temperature does not fall under 15°C while the cooling tower control uses the 21°C set point. Figure 8.14. Transformation of the ambient conditions for the modelled default cooling tower for a specific screw chiller extracted from a manufacturer electronic catalogue (the Milan CAV hourly load curve is used).

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It clearly appears Table 8.19 that the temperature dependency is not linked first to climatic conditions but to the sizing and to the cooling tower choice. It has to be recalled that the cooling towers have been sized for maximum wet bulb temperature with 1% of yearly occurrence. The lower temperature results in Seville as compared to Milan show that on average wet bulb temperatures are higher in Milan. The main differences amongst climates and systems are found for weighting coefficients differences.

Table 8.19. Applying the reduction methodology to the set of available load curves Condenser inlet water temperatures (°C) Weighting coefficients Load 100% 75% 50% 25% 100% 75% 50% 25%

Climate System London CAV 28.3 24.8 21.7 18.5 0.38% 10.96% 41.53% 47.13%

CAV-FC 27.2 26.1 22.3 20.3 2.23% 22.34% 44.54% 30.89% FC4P 28.3 24.5 21.5 18.3 0.32% 10.38% 46.35% 42.96% VAV 26.8 26.2 22.4 18.3 0.07% 9.42% 27.23% 63.28%

Milan CAV 28.8 25.9 21.9 18.4 2.96% 35.53% 39.66% 21.85% CAV-FC 30.4 26.5 23.2 19.8 0.04% 43.26% 42.65% 14.05% FC4P 27.5 25.3 21.6 18.0 1.87% 32.39% 43.29% 22.45% VAV 28.7 26.1 22.5 18.5 2.90% 32.51% 34.75% 29.84%

Seville CAV 29.0 26.2 22.6 18.6 3.31% 38.69% 38.16% 19.85% CAV-FC 29.0 26.2 22.8 19.5 4.15% 45.90% 40.06% 9.89% FC4P 28.9 26.2 22.8 18.6 2.30% 36.20% 38.97% 22.54% VAV 28.9 26.5 23.2 18.8 2.03% 30.28% 42.67% 25.02%

The same market shares can be used for water cooled chillers than for air cooled chillers. because a constant share of water cooled and air cooled systems has been used, the same for all countries. However, since Table 8.19 exhibits different load weighting coefficients, the final weighting for the ESEER for water cooled chillers slightly differs from the air cooled chillers coefficients, Table 8.20.

Table 8.20. Application of the method to derive a ESEER for air cooled chillers used with the free cooling or VAV options

ESEER (Water) Free-cooling VAV

Part load ratio Condenser inlet temperatures (°C)

Weighting coefficients

Condenser inlet temperatures (°C)

Weighting coefficients

Condenser inlet temperatures (°C)

Weighting coefficients

100 28.6 2% 29.4 2% 28.7 2% 75 25.8 34% 26.2 44% 26.2 30% 50 22.3 40% 22.9 41% 22.8 39% 25 18.5 24% 19.6 12% 18.6 29%

8.5 Is there a method good enough for classification of products by order of merit?

We are comparing now the numerical results and the way each of the existing methods would sort chillers by order of merit.

EECCAC final figures -Simplification of the figures and uncertainty estimate Our work clearly shows also that the methodology for air and water cooled chillers enabled to extract seasonal operating temperature conditions with errors on the seasonal efficiencies that are inferior to the experimental uncertainties, for all chillers, included single compressor units. However, it also shows that the experimental uncertainty is quite high. It mainly comes from the uncertainty measurement on the temperature difference at the evaporator. In order to simplify the application of the index, some rounding can be done without modifying noticeably the ESEER figures obtained, largely under the experimental uncertainty. A comparison of the conditions of the 3 available indexes is proposed Table8.21 for air cooled chillers.

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Table 8.21. Comparison of the ESEER conditions with the EMPE and IPLV for air cooled chillers

ESEER ARI EMPE

Part load ratio Temperatures Weighting coefficients Temperatures Weighting

coefficients Temperatures Weighting coefficients

100 35 3% 35 1 % 35 10 % 75 30 33% 26.7 42 % 31.3 30 % 50 25 41% 18.3 45 % 27.5 40 % 25 19 23% 12.8 12 % 23.8 20 %

Temperatures of the ESEER are embedded by EMPE temperatures above and ARI temperature beneath. ESEER weighting coefficients give more weight to the 25% point load than both index. For 50 and 75%, coefficients are nearer to the EMPE index. The 100% coefficient is 3%, nearer from the IPLV one. A comparison of the conditions of the 3 available indexes is proposed Table 8.22 for water cooled chillers.

Table 8.22. Comparison of the ESEER conditions with the EMPE and IPLV indexes for water cooled chillers ESEER ARI EMPE

Part load ratio Temperatures (°C) Weighting coefficients Temperatures Weighting

coefficients Temperatures Weighting coefficients

100 30 3% 29,4 1% 29.4 10%

75 26 33% 23,9 42% 26.9 30%

50 22 41% 18,3 45% 23.5 40%

25 18 23% 18,3 12% 21.9 20%

Temperatures of the ESEER are embedded by the EMPE ones above and ARI temperature beneath except for the 25% point. The ESEER weighting coefficients give more weight to the 25% point load than both index. For 50 and 75%, coefficient are nearer to the EMPE index. The ESEER 100% weighting coefficient is nearer from the IPLV one.

Classification : who is right?

We shall compare now four classifications : according to EER, US-IPLV, EMPE, ESEER, using as a reference the actual EU values obtained by simulation in the three locations and properly weighted. We take the point of view of a user of the Eurovent certification system : by selecting a “better” chiller, am I really selecting a better chiller?

EER is a poor selection tool Figure8.15. Comparison of HSEER with EER on the tested chillers

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IPLV and EMPE are more accurate than EER for classification but do not give enough accuracy for comparison of chillers

Figure 8.16. comparison of US-IPLV with HSEER for the tested chillers

Figure 8.17. comparison of EMPE with HSEER for the tested chillers

HSEER versus EER

0

0,5

1

1,5

2

2,5

3

3,5

4

0 0,5 1 1,5 2 2,5 3 3,5

EER

HSE

ER

HSEER versus IPLV

0

0,5

1

1,5

2

2,5

3

3,5

4

0 0,5 1 1,5 2 2,5 3 3,5 4 4,5

IPLV

HSE

ER

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Based on similar assumptions, the two methods, IPLV and EMPE have the same advantages and disadvantages. The proposed ESEER method allows grading and ranking of chillers by order of merit

Figure 8.18 . comparison of ESEER with HSEER for the tested chillers

Conclusion : the differences are relatively large between existing methods and reality, and not always in the same direction. The newly proposed ESEER method is more accurate in a noticeable manner and satisfies the needs of Eurovent certification process as well as the expectations of the DGTREN in a market transformation effort. First way to realise the testing needed for the ESEER proposed certification method At this point there is still a choice to be made between an experimental approach based on ARI-IPLV (knowing that it will be completely changed in a few years due to the arrival of an ISO standard) or based on the draft CEN standard close to publication and more consistent with the upcoming ISO standard. Since there is no European specific standard to perform part load testing, the analysis is based on :

• the full load testing definition [CEN, 1998], • the IPLV standard [ARI, 1998], that contains some remarks about testing, • the experience gained during the “Joint project” at EDF R&D facility, DMT and manufacturer

laboratories visited.

HSEER versus EMPE

0

0,5

1

1,5

2

2,5

3

3,5

4

0 0,5 1 1,5 2 2,5 3 3,5 4

EMPE

HSE

ER

HSEER versus ESEER

0

0,5

1

1,5

2

2,5

3

3,5

4

0 0,5 1 1,5 2 2,5 3 3,5 4

ESEER

HSE

ER

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The testing must fulfil the following associated constraints : • it must enable to start the part load certification next year, • it must not require the manufacturer presence for the testing, • it must respect the chiller ESEER sequence, • it must minimize the number of testing points, it is to say, the time needed to perform all tests, the

real cost factor. First, if we follow the present ARI approach the manufacturers have to give to Eurovent and to the testing laboratory the expected results close to the ESEER conditions, so as to minimise interpolation and iterations; the computation of ESEER will be only a weighting/interpolation of testing results. In case we want a “blind” checking, more time is needed to guess out the proper conditions for testing in such an ARI approach. We shall present afterwards another way of doing, which seems to us more in the “spirit” of Eurovent certification : once the manufacturer gives the full table of performance, three points are taken “randomly” and checked; it is enough to allow the use of the full table and the calculation of the ESEER. In a certification approach, the manufacturers must give to Eurovent the cooling capacities, the electric powers and the efficiencies of each one of the point that will be tested and the inlet fluid temperature at the condenser according to the ESEER temperature load “curve”. Table 28 gives an example, for chiller number 7, that will be explained just after. Under this format, the table allows to answer directly to the question : what happens to performance when the load (resp. the temperature) decreases.

Table 8.23. For chiller N°7, stage capacities, part load ratio (% of full load, OAT = 35°C), electric power, and efficiencies.

Decreasing capacity

Toe = 7 (°C) Stage 4 Stage 3 Stage 2 Stage 1 Decreasing OAT (°C) % of FL

at 35°C % of FL at 35C % of FL

at 35°C % of FL at 35°C

35 CC 100% 153.7 116.8 81.5 38.4

EP P1 60.0 43.8 26.9 14.2

EER 2.6 2.7 3.0 2.7

30 CC 166.3 82% 126.4 57% 88.1 41.6

EP 53.9 P2a 39.4 P2’ / P2b 24.2 12.7

EER 3.1 3.2 3.6 3.3

25 CC 176.3 134.0 61% 93.4 28% 42.9 EP 48.9 35.8 P3a 22.0 P3’ (P3b) 11.9

EER 3.6 3.7 4.3 Fan cycling 3.6

19 CC 185.0 137.0 93.0 29% 43.9 EP 43.8 32.8 20.7 P4 10.9

EER Fan cycling 4.2 Fan cycling 4.2 Fan cycling 4.5 Fan cycling 4.0

1ST TESTING POINT The testing begins with the full load nominal temperature, the inlet evaporator water temperature is 12°C and the outlet is 7°C. (100%,T1) noted after P1. 2ND TESTING POINT Then, both the outside condensing fluid and the inlet evaporator temperatures are decreased to the 75% load point. It means the inlet water temperature at the evaporator is decreased until 10.75°C. A % tolerance on the cooling capacity is calculated and consequently, the tolerance on the inlet water temperature at the evaporator is known. For continuous control chillers : the chiller can supply the adapted capacity within the allowed tolerance. Point 2, (75%,T2) noted after P2.

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For capacity staged chillers, we guess in general, the chiller cannot supply the required capacity within the allowed tolerance ; thus, either the step is above, either the step is beneath. If a step has been triggered above, it is not counted, and only the step just beneath will be. To reach the step, the inlet evaporator water temperature is decreased until a capacity stage under 10.75°C has been reached. Correlated inlet at the condenser according to the ESEER load versus temperature curve is then imposed, for instance, (70%, T2’) P2’. The procedure enables not to be forced to check the performances of the step above 75%, thus economizing testing time, but at the cost of a further interpolation procedure. Then, for both chillers, with continuous or discontinuous capacities, the following acquisition is made respectively at P2 and P2’ : [ARI, 1998] “C3.1.2 To confirm that steady-state conditions have been established at the specific set of conditions and within the tolerances set forth in C6.2.1, three sets of data shall be taken, at a minimum of five-minute intervals. To minimize the effects of transient conditions, test readings should be taken as nearly simultaneously as possible.” The procedure is then repeated for the 3rd and 4th testing points. For the chiller N°7, the points needed to calculate the ESEER are noted P1, P2’, P3’ and P4. From these points, the interpolation scheme and cycling correction are then applied. The testing time and precision are gathered Table 8.24. Evaluation of the time needed for the complete testing procedure is given with and without the interpolation procedure. In this latter case, two points are needed for the 75% and 50% points. Depending whether two stages embed the 25% load point or not, it will respectively lead to 7 and 6 testing points. The higher option is kept. The points needed at each % load are noted P2a, P2b. It would lead in that case to 6 points. Table 8.24. Evaluation of the first ESEER testing methodology (the +1 testing point corresponds to the nowadays non nominal testing point defined in the Eurovent testing procedure)

4 points with interpolation 4 points without interpolation Testing Time Precision Testing Time Precision

CTS (-) (-) DS (-) (-)

ST [CEN, 1998] 1 hour (+) 1 hour (+)

ST, PID 2 hours (P2’,P3’,P4’) 2 hours

(P2ab,P3ab,P4ab)

FC, PID 1 hour (P3’,P4’) 1 hour

(P3ab,P4ab)

IP WITH : not satisfying WITH : satisfying Testing points 4 (+1) 7 (3 at 1 hour) (+1)

Set up 1 hour 1 hour Disassembling 1 hour 1 hour

TOTAL 2 days 3 days Consequently, the interpolation scheme leads to a non precise enough index for classifying the chillers. But the results can be obtained in 2 days only. When applying the method without interpolation, the ESEER sequence is exact but the testing time increases to 3 days. For both methods, the PID, sensor temperatures and fan cycling problems are likely to create insolvable testing problems. For fan cycling there is no guarantee that the behaviour of the chiller be the same for the tested unit and for a sold unit, since the parameter could be modified to get favourable behaviour for testing. The same rationale applies to chillers that have several possible staging to output the same capacity.

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These two formats of the same scenario are not satisfying ; as a consequence, the second scenario is proposed. Second way to realise the testing needed for the ESEER proposed certification method The manufacturer must give the Table 8.25 to Eurovent for the chiller that will be tested. All the stages programmed in the soft of the chiller must be supplied. For continuous control chillers, 6 capacity steps at least must be supplied. Table 32 is the Eurovent tested chiller N°7 : a 2 circuits, 4 scroll compressors (1 tandem by circuit), 3 fans on each circuit.

• C1 and C2 are the 2 distinct circuit. • The percentages refer to the full load point at 35°C OAT for the cooling capacity, the electric power

and the EER. • The percentages for fans and compressors refers to the circuit full load and not to the chiller full

load. Here they are electric power ratios. For chiller specific configurations, supplementary information should be gathered for testing :

• For fans, their position should be given to the experimenter if they supply the air for different part of the air coil and that consequently stopping 1 fan is not equivalent to stopping another one. For variable speed chillers, the manufacturer should also explain to the experimenter how to reach the published points.

• For reciprocating chillers, supplementary information should be supplied to the experimenter to enable to make the difference between compressor unloading or compressor ON-OFF.

• For screw chillers with slide valve, the manufacturer should explain to the experimenter how to activate the slide valve (access and postion).

Table 8.25. Scenario 2, part load and reduced temperature performance table

Toe = 7 (°C) Stage 1 Stage 2 Stage 3 Stage 4

OAT (°C)

% (full load

35°C)

% (full load

35°C) % (full load

35°C) % (full load 35°C)

19 C1 C2 CC 29% 43.9 C1 C2 CC 61% 93.0 C1 C2 CC 89% 137.0 C1 C2 CC 120% 185.0

Fan 66% 0% EP 18% 10.9 66% 66% EP 35% 20.7 66% 66% EP 55% 32.8 66% 66% EP 73% 43.8

Comp 50% 0% EER 157% 4.0 50% 25% EER 175% 4.5 50% 100% EER 163% 4.2 100% 100% EER 165% 4.2

25 C1 C2 CC 28% 42.9 C1 C2 CC 61% 93.4 C1 C2 CC 87% 134.0 C1 C2 CC 115% 176.3

Fan 66% 0% EP 20% 11.9 100% 100% EP 37% 22.0 100% 100% EP 60% 35.8 100% 100% EP 82% 48.9

Comp 50% 0% EER 141% 3.6 50% 25% EER 166% 4.3 50% 100% EER 146% 3.7 100% 100% EER 141% 3.6

30 C1 C2 CC 27% 41.6 C1 C2 CC 57% 88.1 C1 C2 CC 82% 126.4 C1 C2 CC 108% 166.3

Fan 100% 0% EP 21% 12.7 100% 100% EP 40% 24.2 100% 100% EP 66% 39.4 100% 100% EP 90% 53.9

Comp 50% 0% EER 128% 3.3 50% 25% EER 142% 3.6 50% 100% EER 125% 3.2 100% 100% EER 120% 3.1

35 C1 C2 CC 25% 38.4 C1 C2 CC 53% 81.5 C1 C2 CC 76% 116.8 C1 C2 CC 100% 153.7

Fan 50% 0% EP 24% 14.2 100% 100% EP 45% 26.9 100% 100% EP 73% 43.8 100% 100% EP 100% 60.0

Comp 25% 0% EER 106% 2.7 50% 25% EER 118% 3.0 50% 100% EER 104% 2.7 100% 100% EER 100% 2.6 Three points are tested on the whole map. Indeed, since the complete performance map is known, it is not needed any longer to test the ESEER specific points to be able to calculate the index value. As a consequence, 3 points (+1 for the non nominal temperatures full load point) only can be chosen randomly by the experimenter. Then the ESEER can be calculated following the scheme:

• 25%, 50% and 75% load point efficiencies are calculated according to equation (7a) and (8a), • If needed, cycling correction is done according to Equation (2) with Ccyc= 0.9.

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Using this scenario, most of the testing problems are solved .It still remains to manufacturers that build chillers with variable speed fan control to make sure a precise fan speed reduction can be set manually. As for the precedent scenario, it is not guaranteed the sold chillers will have exactly the same characteristics. But this problem is part of the Eurovent certification scheme.

Table 8.26. Evaluation of the second ESEER testing methodology 4 points with interpolation

Testing Time Precision CTS NO DS NO

ST [CEN, 1998] 1 hour (+) ST, PID NO FC, PID NO

IP (+) Testing points 3 (+1)

Set up 1 hour Disassembling 1 hour

TOTAL 1 day Final choice of the ESEER testing methodology This testing procedure highly reduces the testing time as compared to the scenario 1. Moreover, it will enable to easily set up other seasonal performance indexes that the average ESEER. If the performance map (Table 8.25) is published, it will be a huge and needed progress :

• it will allow the buyer to compare the chillers on specific site conditions ; at the moment only the EER information is available, and we have seen it contained in fact little average efficiency information,

• it will also enable the buyer to optimise the chiller size as a function of the specified site load curve. At the knowledge of the authors, 4 manufacturers on the European market have already achieved similar to Table 8.25 performance maps for each chiller. Only the fan and compressor information have to be added to enable the testing. Waiting that performance maps be ready for all manufacturers, the certification procedure can begin with only the information that enable to characterize the ESEER points, it is to say the highlighted testing points Table 8.25. Perspective of the proposed ESEER The tool is not gifted of any prediction power of the yearly consumption for any real installation. It is just an indicator of the seasonal performance, whose only aim is to classify the chillers a fairer way the simple EER does. It has been shown that the reduction methodology enabled to successfully extract four weighted temperature conditions : the bias introduced was lower than the experimental uncertainty. As a consequence, it could be applied to any other stock of load curves to build other indexes for other domains. This method also enables to increase the number of points ; however, for some of the load curves considered, more than 5 points would give a useless 100% load point. Given the differences within the weighting coefficients for the 3 selected climates, different values should be published at least as a function of the country. To do so, a simple method based on cooling degree days or more simulations could be developed, if manufacturers require it, to adapt the coefficients and temperatures by country. Similar spreading could also be done by type of building. Moreover, the present study enabled to

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show that the free cooling and VAV options should be differentiated from typical CAV installations, mainly because of the differences in the weighting coefficients. As given, the coefficients are nearer from Southern Europe operating temperatures and weightings (Italy, Spain and Greece represent 85% of the installed chiller based systems according to the EECCAC stock and market study). The load weighting coefficients are the main seasonal efficiency drivers. Temperatures can be shifted easily of 1°C if needed (as practised to round the operating temperatures). This index could be used for single circuit units. However, a method must be adapted to determine the cycling degradation versus the load. The default cycling law could then be revised for all chillers. The reduction methodology for a dedicated load curve associated to the presentation of individual testing results or more generally of part load performances and reduced temperature efficiencies for chillers would be a highly efficient simple selecting tool for buyers when following the choice method steps hereafter :

• Hourly simulation of the project gives the building or chiller hourly load curve for the specific project.

• Then, the reduction methodology enables to characterize 4 capacity step points and operating temperatures.

• The presentation of part load results by the manufacturers enables to select the chiller on 4 efficiency points.

Certifying seasonal performances for chillers means indicating an average efficiency generally higher than nominal efficiency as has been largely figured. But it also means to avoid the efficiency competition may be based on non-representative indicators, as nominal full load EER is. Thus, the seasonal performance index is thought to be a huge and necessary progress to strengthen energy efficiency of chillers. Future versions should consider the extension to single stage units that generally operate at different conditions. It has been shown however that the methodology could be applied to these units for the load curves treated. Similar work has to be performed in the heating mode since air to water reversible chillers is an increasing end use in Europe. The same philosophy could also be applied to ground water condensing chillers and heat pump, also a developing market in France and Germany. The applicability to each country and building type should also be studied in order to give a full range of testing conditions and weighting coefficients nearer from specific and climatic applications. In order to approach field reality, the integration of dynamics into the part load testing index should also be considered. Nevertheless, supplementary work has to be performed to reach a such far away goal from actual chillers characterization. This work could also serve as a basis for developing a seasonal index for room air conditioners, the more developing end-use in Europe nowadays.

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9. POLICY OPTIONS AND RECOMMENDATIONS TO IMPROVE CAC ENERGY PERFORMANCE 9.1 Some fundamental considerations regarding policy measures The results of the energy scenario analyses of the preceding chapters have illustrated that there are significant potential energy savings to be attained by the optimisation of CAC systems. As CAC systems are diverse in nature, are often designed on-site rather than simply being factory made packaged-systems and are installed and operated in diverse circumstances, viable policy measures will need to take account of the diverse circumstances which apply to them in order to realise the potential energy savings. CAC equipment, like other tradable goods are subject to the terms of the European single market and therefore it is appropriate for the European Commission to develop policy measures which will raise the average efficiency of new equipment sold within the European Union. These type of measures include certification, energy labelling, and minimum energy performance requirements (either mandatory or voluntary in nature). Proposals for each of these are made in section 9.3. In the case of minimum energy performance requirements, these could be developed within the mandate of the proposed Directive on Ecodesign of End-Use Equipment; however, the application of energy labelling for central air conditioning equipment would either require an amendment of the existing energy labelling Directive to include energy-using equipment destined for usage sectors other than just households, or it would require the issue of a new primary Directive giving authority to the Commission to develop energy labels for commercial and tertiary equipment.

Important as these measures are, they only address the efficiency of the end-use equipment used in CAC systems as determined under standard test conditions and will not realise many of the potential energy savings which are related to the design, installation and operation of specific CAC systems. Policy measures which can realise these savings at the design and installation stage are typically provided through building thermal regulations. The new Energy Performance in Buildings Directive places some obligations on Member States to develop policy measures which will address some part only of these savings; however, there are many more areas for action than are specified within it. The most advanced national building thermal regulations addressing the energy consumption of central air conditioning systems are in the UK and Portugal; yet even these are not as mature or as encompassing as the US ASHRAE 90.1-1999 standard. Many Member States are in the process of revising their building regulations to take account of the requirements of the Energy Performance in Buildings Directive and thus this offers an ideal opportunity for them to co-operate at least regarding measures applying to the energy performance of building cooling systems. Specific proposals regarding this are made in sections 9.2, 4 and 5.

9.2 Policies and measures to encourage the selection of more efficient equipment The analysis presented in this study has shown that there is a significant variation in energy efficiency for all types of CAC equipment that have been investigated when tested under standard test conditions. The measures which can be considered to encourage a higher energy efficiency levels for new CAC equipment are:

• Provision of information (labelling, grading, efficiency ratings)

• Removing less efficient models from the market (MEPS or voluntary agreements)

• Encouraging higher sales-weighed average efficiency levels through negotiated agreements (e.g. fleet-average efficiency targets)

• Financial and/or fiscal incentives for higher efficiency equipment

• Public procurement and general market transformation programmes

Measures to provide information to end-users and equipment procurers At the current time there are no mandatory requirements to provide end-users with information on the energy efficiency of central air conditioning equipment. Manufacturers generally test the efficiency of their equipment and report the results in their product literature and catalogues. Without independent verification

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these values may lack credibility therefore Eurovent currently operates a certification scheme of their members products in order to ensure the reliability of claimed performance levels. While Eurovent Certification reports the energy efficiency levels of most CAC equipment as tested under standard test conditions it does not yet compare the efficiency of equipment in a simplified form, thus the user of the information is required to have a high degree of expertise to be able to interpret the reported efficiency level. An effective way of indicating the relative energy efficiency of products in a simplified manner is to use a categorical energy label, i.e. to apply a label which rates the energy performance of the equipment into one of a limited set of efficiency ranges which are graded using a simple scale. The current EU energy label applied to household appliances, which grades appliance energy efficiency using an A to G rating scale, is an effective and widely known example of this.

CAC equipment are generally large and are not generally ‘on display’ in a shop window at the moment of purchase. It is therefore debatable whether there is much advantage in applying a removable energy label to the product itself; however, ensuring that the purchaser and subsequent end-user be able to see the relative efficiency of the equipment is likely to be an effective market transformation tool. The information provided in a comparative energy label provides the basic language of energy efficiency that enables many other market transformation efforts to be realised. Even though central AC equipment are subject to the classic split incentives situation where the purchaser or procurer is unlikely to be the entity responsible for paying the energy bills the provision of relative efficiency information is still a fundamental component supporting more complicated policy measures that may aim to bridge the split incentives problem. It is therefore strongly recommended that mechanisms be put in place which will allow such information to be passed through the equipment procurement and usage chain.

The current EU energy labelling Directive is restricted to household appliances hence would require amendment to address this issue. It certainly makes sense to exploit the high brand recognition and public understanding of the current A to G energy-label efficiency scale for the comparative rating of CAC components. However, a key question is how that information should be presented to the public? In the case of products having split-incentives such as CAC components do, it is desirable that not only the procurer should be aware of the comparative energy efficiency of the equipment they are purchasing but also that subsequent potential users of the piece of equipment should be able to see this easily. As mentioned a removable paper label as currently applied to household appliances is not likely to satisfy these requirements. Alternatively it seems essential that the information on the comparative energy rating of the equipment (A to G) should be presented in all product catalogues and literature, including on-line sales, and that the label information should be indicated on the fixed metal rating plate that is applied to the equipment before it leaves the factory.

For the time being there is no such scheme in place and it may be some time before one is formally developed; however, industry associations, such as Eurovent, have expressed an interest in adopting such a comparative grading approach, which could be applied by them on a voluntary basis (e.g.. it could be made mandatory within Eurovent for all manufacturers who wish to place their products in the Eurovent catalogue to report the A to G grading of their equipment). By so doing it would allow manufacturers to express the relative energy performance of their products and, as has been seen for household appliances, would allow greater differentiation of products within the market place. The European Commission could take advantage of this informal adoption of a grading system to prepare the ground for a formal efficiency rating scheme in the years ahead. In the spirit of aiding the rapid adoption of an efficiency grading scheme for central air conditioning equipment, the remainder of this section contains explicit recommendations regarding the thresholds which could be applied to denote the A to G efficiency grades for each equipment type examined in this study.

A to G efficiency grading of central air conditioner components This section gives proposals for the energy efficiency grading of the principal types of CAC cooling equipment (the equipment which generates the primary cooled or heated fluid) separated from the rest of the CAC system (the cold or heat transfer and conditioned air distribution system) under standardised load conditions. The grading follows the A to G approach used in the EU energy label for household appliances and also applied in some national regulations to rate the energy performance of buildings and cars.

71

Structural issues

There are certain technical structural issues that need be addressed prior to the formulation of proposals for an A to G efficiency grading.

Full or part-load? The first issue to consider is whether the scale should be based on full or partial load ratings. It is clear from the results reported in Chapter 8 that a part-load efficiency scale is more representative of the true energy efficiency of CAC components when in real usage conditions and therefore the ultimate goal should be to develop ratings based on part-load performance; however, at the present time there is no accepted part-load test and only full load efficiency ratings are generally available. The analysis of Chapter 8 has led to a proposal for an EU IPLV for chillers based upon specific rating conditions; however, even if this is adopted without controversy and in the shortest imaginable time scale it is still likely to be years before there are a large number of EU IPLV ratings available from which to develop an IPLV rating scale. Thus a more pragmatic solution would be to adopt an efficiency grading scale based on an analysis of the currently available full-load efficiency ratings and to use this in the near term. The results of Chapter 8 have indicated that commonly, but not always, the relative efficiency of equipment determined at full load is indicative of its relatively efficiency at part-load, therefore there is little risk of misleading the public by adopting an initial grading system based on full load performance. Furthermore, the part-load ratings are usually equivalent to or slightly higher than the full load ratings, which suggests that the same efficiency range may be applicable to both the full and part-load efficiency grades. Accordingly this report presents proposals for A to G efficiency grades based on the analysis of full-load performance data, which are intended for use in the near term. In the future, at such a time when EU part-load ratings are widely available, it would be appropriate to review the appropriateness of these grades for translation into part-load grades.

Heating and cooling-modes. Two separate efficiency scales letters (one for the cooling function and one for the heating function) are already used for RAC in the EU RAC energy labelling Directive. Therefore the same approach is followed here.

Product categories. A key question is whether it is appropriate to mix all the chillers into a single product group for efficiency grading or to adjust the efficiency scales depending on the product sub-category? Were the systems using water and air completely comparable (i.e. were the energy consumed by the cooling tower to be included) it would be possible to use the same scale for both; however, the uncertainties regarding the tower control and the origin of the natural water being used are such that it is impossible to make a meaningful comparison.

The basic chiller types to be separated correspond at least to the different testing conditions applied in the test standard, which inherently generate incomparable figures as follows:

• water cooled (in cooling and/or heating-mode for reversible units),

• air cooled (in cooling and/or heating-mode for reversible units),

• floor-feed systems (in cooling and/or heating-mode for reversible units);

• condenser-less units.

The efficiency ratings are not directly comparable between these 7 distinct sets of test conditions.

The chillers could also be separated into two categories as a function of cooling capacity say, those with a capacity less than 750kW and those with a capacity greater than 750kW. The separation for units currently in the catalogue is not needed since the screw units (mostly) follow the same design as units with less than 750 kW of cooling capacity and the product ranges overlap across both capacity ranges. Some centrifugal units are already integrated in the catalogue; although it seems they can be graded on the same scale.

Some markets require ducted condensers, which degrade the EER. It is proposed to introduce a specific classification for them. Admittedly the additional consumption of the necessary fan for the condenser is not included according to the test standard but this omission is arbitrary (efficiency of 0.3) and in fact heat

72

exchange drops in this case, and the correction of the standards cannot translate the reality. Establishing a classification based on specific statistics for ducted units will eliminate the two problems. However, when testing ducted units, the available static pressure is set by the design of the manufacturer. For this reason, the testing does not give comparable figures. Therefore it is recommended that manufacturers should supply a common static pressure for testing ducted units.

Proposed grading scale

In order to require the same effort for each type of chiller, it is proposed to make the average efficiency level of the products on the market correspond to the threshold between the D and E grades for each chiller type whenever possible. The average EER and COP values are of course based on data which include units using HFCs. The following classes and values regarding R22 are given as an information only since the European market will no longer have R-22 chillers.

The construction of the scale for the different categories, intends to respect the following rules, classified by order of importance:

• use of the same classes width (for simplicity),

• use of limits of classes ending by 0.05 or 0.1,

• adjustment of the extremes (G of about 10 %, A of about 1 %),

• centering on the average of the catalogue (equal treatment between types),

From these basic rules come the following proposals (Tables 9.1 and 9.2).

Table 9.1 Proposed efficiency grades for chillers in the cooling-mode

EER boarders Air Cooled Air Cooled,

Floor heating and cooling

Water Cooled Water Cooled, Floor heating and cooling

Remote condenser

A/B 3.10 3.65 5.05 4.75 3.55

B/C 2.90 3.50 4.65 4.60 3.40

C/D 2.70 3.35 4.25 4.45 3.25

D/E 2.50 3.20 3.85 4.30 3.10

E/F 2.30 3.05 3.45 4.15 2.95

F/G 2.10 2.90 3.05 4.00 2.80 Note: for borders, A, for air-cooled units, is strictly superior to 3.10.

Table 9.2 Proposed efficiency grades for chillers in the heating-mode

COP boarders Air Cooled Air Cooled,

Floor

Water Cooled Water Cooled, Floor

A/B 3.25 4.20 4.45 4.50

B/C 3.05 4.05 4.15 4.25

C/D 2.85 3.90 3.85 4.00

D/E 2.65 3.75 3.55 3.75

E/F 2.45 3.60 3.25 3.50

F/G 2.25 3.45 2.95 3.25 Note: for borders, A, for air-cooled units, is strictly superior to 3.25. Impact of the proposed grading on the chiller cooling market

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Figure 9.1. Air cooled chillers, cooling-mode, < 750kW

0.0%

5.0%

10.0%

15.0%

20.0%

25.0%

30.0%

35.0%

40.0%

A (>3,1) B (>2,9) C (>2,7) D (>2,5) E (>2,3) F (>2,1) G (<2,1)

R407CR134aR22HFC

refrigerant R407C 2.41 refrigerant R134a 2.55 refrigerant HFC 2.42 refrigerant R22 2.59

A (>3,1) B (>2,9) C (>2,7) D (>2,5) E (>2,3) F (>2,1) G (<2,1) Total

SUM R407C 4 30 104 296 411 310 121 1276 % R407C 0% 2% 8% 23% 32% 24% 9% 100%

SUM R134a 0 11 39 27 30 27 4 138 % R134a 0.0% 8.0% 28.3% 19.6% 21.7% 19.6% 2.9% 100.0% SUM R22 14 20 118 163 108 33 30 486

% R22 3% 4% 24% 34% 22% 7% 6% 100% %HFC 2% 5% 25% 30% 22% 10% 5% 100%

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Figure 9.2. Water cooled chillers, cooling-mode, < 750kW

0%

5%

10%

15%

20%

25%

30%

35%

40%

45%

A (>5,05) B (>4,65) C (>4,25) D (>3,85) E (>3,45) F (>3,05) G (<3,05)

R407CR134aR22HFC

refrigerant R407C 3.79 refrigerant R134a 4.34 refrigerant HFC 3.85 refrigerant R22 3.88

A (>5,05) B (>4,65) C (>4,25) D (>3,85) E (>3,45) F (>3,05) G (<3,05) Total SUM R407C 0 31 35 26 45 87 17 241 % R407C 0% 13% 15% 11% 19% 36% 7% 100% SUM R134a 0 6 11 7 4 0 0 28 % R134a 0.0% 21.4% 39.3% 25.0% 14.3% 0.0% 0.0% 100.0% SUM R22 5 3 15 19 28 14 10 94 % R22 5.3% 3.2% 16.0% 20.2% 29.8% 14.9% 10.6% 100.0% %HFC 0% 14% 17% 12% 18% 32% 6% 100%

When we compare the model-based statistics from the directory with a sample of 2001 confidential market-based figures we find the values given in Figures 9.3 and 9.4.

75

Figure 9.3. Sales and Market statistics

Sale statistics R407C

0

500

1000

1500

20001,6

1,71,8

1,9

2

2,1

2,2

2,32,4

2,52,6

2,7

2,8

2,9

3

3,1

3,23,3

Eurovent catalog values R407C

050

100150200250

1,61,7

1,8

1,9

2

2,1

2,2

2,32,4

2,52,6

2,7

2,8

2,9

3

3,1

3,23,3

Sales and Eurovent catalogue statistics for air condensing units are convergent: sales statistics show a better efficiency.

Figure 9.4. Sales and Market statistics

Sale statistics R407C

0100200300400500

2,72,9

3,1

3,3

3,5

3,7

3,94,14,3

4,5

4,7

4,9

5,1

5,3

5,5

Eurovent catalog values R407C

0

5

10

15

202,7

2,9

3,1

3,3

3,5

3,7

3,94,14,3

4,5

4,7

4,9

5,1

5,3

5,5

Sales and Eurovent catalogue statistics for water condensing units are convergent: sales statistics show a better efficiency..

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Proposal of grading of packaged AC in Europe (extension of the RAC labelling scheme)

(1) Air-cooled air conditioners - cooling mode

Table 9.3 Proposed efficiency grades for large-split packaged AC in the cooling-mode

Energy efficiency class Energy efficiency ratio

A 3.20 < EER

B 3.20 > EER > 3.00

C 3.00 > EER > 2.80

D 2.80 > EER > 2.60

E 2.60 > EER > 2.40

F 2.40 > EER > 2.20

G 2.20 > EER

Table 9.4 Proposed efficiency grades for large unitary packaged AC in the cooling-mode

Energy efficiency class Energy efficiency ratio

A 3.00 < EER

B 3.00 > EER > 2.80

C 2.80 > EER > 2.60

D 2.60 > EER > 2.40

E 2.40 > EER > 2.20

F 2.20 > EER > 2.00

G 2.00 > EER

(2) Water-cooled air conditioners – cooling-mode

Table 9.5 Proposed efficiency grades for split and multi-split packaged AC in the cooling-mode

Energy efficiency class Energy efficiency ratio

A 3.60 < EER

B 3.60 > EER > 3.30

C 3.30 > EER > 3.10

D 3.10 > EER > 2.80

E 2.80 > EER > 2.50

F 2.50 > EER > 2.20

G 2.20 > EER

Table 9.6 Proposed efficiency grades for unitary packaged AC in the cooling-mode

Energy efficiency class Energy efficiency ratio

A 4.40 < EER

B 4.40 > EER > 4.10

C 4.10 > EER > 3.80

D 3.80 > EER > 3.50

E 3.50 > EER > 3.20

F 3.20 > EER > 2.90

G 2.90 > EER

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(3) Air-cooled air conditioners – heating mode

Table 9.7 Proposed efficiency grades for large-split packaged AC in the heating-mode

Energy efficiency class Coefficient of performance

A 3.60 < COP

B 3.60 > COP > 3.40

C 3.40 > COP > 3.20

D 3.20 > COP > 2.80

E 2.80 > COP > 2.60

F 2.60 > COP > 2.40

G 2.40 > COP

Table 9.8 Proposed efficiency grades for large unitary packaged AC in the heating-mode

Energy efficiency class Coefficient of performance

A 3.40 < COP

B 3.40 > COP > 3.20

C 3.20 > COP > 3.00

D 3.00 > COP > 2.60

E 2.60 > COP > 2.40

F 2.40 > COP > 2.20

G 2.20 > COP

(4) Water-cooled air conditioners – heating mode

Table 9.9 Proposed efficiency grades for split and multi-split packaged AC in the heating-mode

Energy efficiency class Coefficient of performance

A 4.00 < COP

B 4.00 > COP > 3.70

C 3.70 > COP > 3.40

D 3.40 > COP > 3.10

E 3.10 > COP > 2.80

F 2.80 > COP > 2.50

G 2.50 > COP

Table 9.10 Proposed efficiency grades for unitary packaged AC in the heating-mode

Energy efficiency class Energy efficiency ratio

A 4.70 < COP

B 4.70 > COP > 4.40

C 4.40 >COP > 4.10

D 4.10 > COP > 3.80

E 3.80 > COP > 3.50

F 3.50 > COP > 3.20

G 3.20 > COP

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Market mixed statistics based on the scheme (splits and packages mixed)

Table 9.11 Air-cooled air conditioners – cooling-mode -mixed statistics (splits and packages mixed). Market average efficiency = 2.46 W/W.

Class definition % on market Grade % with equal class width

3.20 < EER 2% A 2%

3.20 > EER > 3.00 5% B 5%

3.00 > EER > 2.80 7% C 7%

2.80 > EER > 2.60 15% D 15%

2.60 > EER > 2.40 22% E 22%

2.40 > EER > 2.20 26% F 26%

2.20 > EER 23% G 11%

Figure 9.5. Air-cooled air conditioners - Cooling function -mixed statistics

% of market(density)

2%

5%

7%

15%

22%

26%

11%

0%

5%

10%

15%

20%

25%

30%

A B C D E F G

Table 9.12 Air-cooled air conditioners - Heating function -mixed statistics (splits and packages mixed). Market average efficiency = 2.87 W/W.

Class definition % on market Grade % with equal class width

3.60 < COP 3% A

3.60 > COP > 3.40 5% B 5%

3.40 > COP > 3.20 6% C 6%

3.20 > COP > 2.80 46% D 23%

2.80 > COP > 2.60 17% E 17%

2.60 > COP > 2.40 11% F 11%

2.40 > COP 8% G 5%

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Figure 9.6. Air-cooled air conditioners - Heating function -mixed statistics

% of market(density)

0

0,05

0,1

0,15

0,2

0,25

A B C D E F G

The last column gives a frequency independent of class width (so divided by two in the intermediate class for heating. It's a pity that our statistics mix the various subtypes (packages and splits) because it seems that there are really two populations in the data. Note also that we have here only full load COP and EER.

Removing less efficient equipment from the market (MEPS and voluntary agreements) The draft Ecodesign of End-Use Equipment Directive proposes that policy measures should be enacted which bring the market average efficiency of equipment up to the least-life cycle cost for the final user. This implies the introduction of energy efficiency requirements for new equipment, which could be mandatory (MEPS) or voluntary in nature but attaining the same goal. Table 9.13 lists the full-load efficiency levels for CAC equipment associated with the least life cycle cost for the final user as determined in this study. The adoption of policy measures which would move the average new equipment efficiency levels to those indicated in Table 9.3 from 2005 onwards was simulated in Scenario 2 as reported in Chapter 7 and would lead to energy savings of about 11% by 2020 compared with the Business As Usual scenario. Adopting similar measures based upon an EU part-load performance indicator (EU IPLV and/or EU SEER) would produce energy savings in 2020 at about 18% of the Business As Usual scenario total. By contrast simply setting MEPS at the current average full-load efficiency levels for CAC equipment would only save energy equivalent to 3% of all CAC energy consumption by 2020: see Scenario 1 in Chapter 7.

Table 9.13 Full-load efficiency levels associated with the least-life cycle cost by CAC component (W/W).

Equipment type Cooling capacity range

(kW)

EER at least life cycle cost

(W/W)

Water-cooled chillers standard 12 to 750 4.50

Water-cooled chillers centrifugals 750 upwards 5-6

Air-cooled chillers 12 to 750 3.00

Large packaged AC (cooling mode) 12 to 75 3.22

RAC 3 to 12 3.20

An alternative approach to MEPS setting is to harmonise levels with those applied internationally. From a commercial perspective there is some logic to this approach, because most companies supplying the EU

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market are multinationals with headquarters outside the EU and the CAC equipment they supply to the EU market is based on the same product platforms that are used in other international markets. From a programme management perspective this has the advantage that the MEPS efficiency thresholds are tried and tested having been successfully applied elsewhere. Furthermore in the only case where the life cycle cost analyses have been directly compared, for packaged AC units, the least life cycle cost efficiency level has been found to be the same, which suggests that the US efficiency thresholds might be appropriate for adoption in the EU.

Were the EU to adopt the same MEPS requirements as currently apply in the US through ASHRAE 90.1-1999 (as either a mandatory or voluntary measure) it would imply the following efficiency thresholds:

Table 9.14 ASHRAE 90.1 MEPS levels for CAC equipment (W/W)

Equipment Type MEPS level (W/W) Capacity range

Packaged AC (cooling-only) 3.02 19.5 to 39.5

Packaged AC (cooling-only) 2.84 39.5 to 70.3

Packaged AC (reversible) 3.02 19.5 to 39.5

Packaged AC (reversible) 2.84 39.5 to 70.3

Water-source heat pump, 4tons (cooling-mode)

3.51 Values about 14.1

Water-cooled screw chiller, 125tons

4.45 COP (4.50 US IPLV) Values about 440

Water-cooled centrifugal chiller, 300tons

6.10 COP (6.10 US IPLV) Values about 1056

Of course there are other options than applying a mandatory minimum energy performance threshold, which could achieve similar objectives. One approach would be to negotiate fleet-average efficiency targets with the European industry. This would have the benefit that it would not have to wait that the Ecodesign of End-Use Equipment Directive be implemented.

Encouraging the selective acquisition of more efficient equipment by other means There are many actions that Member States can initiate to encourage the selective procurement of more efficient central air conditioning equipment, which could take advantage of any efficiency grading system that is introduced. Some of potential measures are as follows:

• Establish public sector procurement guidelines (e.g. only class A equipment should be procured for use in the public sector)

• Develop corporate procurement guidance documentation, analytical tools and training materials to explain and quantify the advantages of procuring more energy efficient CAC equipment

• Develop and promote on-line directories of efficient equipment (e.g. as in the UK web-site www.ukepic.org)

• Develop low cost credit lines for more efficient equipment (e.g. as in the UK Enhanced Capital Allowance scheme)

• Arrange training programmes for associations of designers and installers to explain the cost-benefits of more efficient AC equipment and to ensure lines of access to efficient equipment

• Provide rebates on efficient equipment (e.g. as in the Dutch rebate scheme for class A household appliances)

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• Create favourable tax differentials for efficient equipment (e.g. lower VAT levels, or corporate tax breaks for manufacturers and/or corporate procurers)

9.3 Policies and measures to encourage the adoption of more efficient system structures

Policy aims and potential measures targeting the adoption of more efficient system structures Measures that aim to encourage the adoption of more efficient components, as outlined in section 9.2, will only realise some of the potential to save energy for CAC systems. Such measures are necessarily focused on the individual components and not on the performance of the system as a whole, therefore they do not encompass the freedoms and constraints applying to the system designer in trying to design an efficient yet effective CAC system. Furthermore they do not address the activities of the installation engineer who is responsible for executing the system design and commissioning the system. The comparison of the energy performance of eighteen different CAC systems, each designed to provide total cooling in a typical EU office, has shown that there can be a difference of a factor of 2 in the total energy consumption per m2 of cooled space for typical configurations of CAC equipment using average efficiency components. The same results have also indicated that while the proportion of energy required for heating and cooling may vary appreciably from one climate and Member State to another, the absolute annual energy consumption per m2 shows a much smaller variation and follows a trend that can be related in a roughly proportional manner to the annual cooling and heating degree days. These limited results imply that it might be feasible to develop simple benchmarks of overall CAC system performance as a function of the level of cooling and air quality required, the building type and of the cooling and heating degree days.

Policy measures would aim to encourage the adoption of more efficient CAC systems while maintaining the freedom of the system designer to achieve a solution which meets the cooling, environmental and air quality requirements of the client within acceptable cost boundaries.

As such building codes are the primary policy measure which promote the adoption of more efficient system types; however, these can be supported by the following measures:

• The provision of analytical tools and technical guidance enabling the energy efficiency of CAC systems to be optimised

• Training of system designers and installers on the options regarding energy efficient CAC systems

• The provision of financial and fiscal incentives to help overcome split incentives such as the provision of cheap credit for efficient systems

Building code requirements are articulated in quite different ways among EU Member States. One difference regards how the requirements for energy using systems in building codes should be expressed. In the UK the primary policy goal is carbon abatement and therefore the requirements are expressed in terms of maximum allowable emissions of carbon per m2 per year. In some Member States the building codes are articulated in terms of limits regarding the maximum allowable energy consumption per m2. Once the fuel of the heating and cooling system has been fixed these two approaches are effectively transposable; however, the carbon approach provides an additional degree of freedom for designers to satisfy the requirements through optimisation of the choice of fuel used by the system. The approach of setting limits for either energy consumption or carbon emissions per m2 leaves the designer almost complete freedom to decide how they are going to satisfy the requirements. In Portugal, however, there are no requirements on overall annual energy or carbon but rather a set of simple prescriptions to follow, which are designed to save energy. Both approaches have their merits and indeed can even be integrated as is the case in the US ASHRAE 90.1-1999 standard. In this standard the designer is obliged to follow some general requirements and mandatory measures for each technical section but thereafter has a choice of two final compliance pathways: following a further set of simple prescriptive requirements or demonstrating compliance by satisfying the “energy cost budget”, method which requires the use of one of a number of designated detailed building thermal simulation tools. Following the simple prescriptive measures is an easy way for a designer to demonstrate

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their compliance with the standard; however, the prescriptions are relatively rigorous compared with the requirements when a detailed simulation tool is used. Therefore the standard creates an incentive for designers to use detailed building thermal simulation software. The combination of a prescriptive compliance pathway and a pathway based on meeting overall energy limits demonstrated, through the use of detailed simulation software, simultaneously meets the needs of designers “in a hurry” dealing with standard design briefs and those who have specific and complicated design briefs; who may need more freedom to meet the same energy goals than would otherwise be allowed through application of a set of simple prescriptions.

Legal basis for policy measures targeting more efficient system structures As mentioned, building codes are the primary policy measure available to encourage the adoption of more efficient system types. The recent Energy Performance in Buildings Directive obliges Member States to develop mandatory minimum energy performance requirements for buildings and specifies that these should encompass the energy used by mechanical cooling systems. In the case of the UK this Directive may simply require minor revisions of the existing regulations for building cooling energy performance, but for most other Member States it will require completely new regulations to be developed addressing the cooling system. The requirement that a simple calculation method should be constructed against which the compliance of the MEPS is to be judged may also require modification of the Portuguese regulations. The need to develop minimum energy performance requirements within building codes for so many EU states raises the question of whether it would be appropriate for Member States to co-operate with each other. The ENPER-TEBUC study within SAVE deals with the issue of harmonisation in European Building Codes and has set up a platform for information exchange on Energy Performance (EP) standardisation and legislation among the prominent national players. The intention has been to systematically collect and summarise the different approaches and to develop suggestions for a European 'model code'. Ultimately such a code could be the EU equivalent to the US ASHRAE 90.1 standard, which is non-binding in itself but can be brought either wholly or in-part into national regulations as deemed appropriate by the authorities in each Member State. Sharing development and analytical effort makes considerable sense for such a major undertaking and is a key recommendation from this study. In the longer term the EU model code could be designed to enable a energy efficiency labelling or grading system for CAC systems of say an A to G type. The difficulty in obtaining the best grades (closer to A) should be increasing and involve not only the manufacturer but also other elements of the chain. Moving from G to F or E might be based only on full-load ratings i.e. readily available EER values. Higher grades such as E or D could be reached on the basis of a certain value of SEER, taking part load optimisation efforts into account. Part load is not only a phenomenon to be computed, it's by itself one of the most important energy saving features. Following these measures the importance of system design cannot be neglected. A designer could refer to a design procedure proving he has considered all cost effective options of the project. and reached a certain performance level like C (- 25 %) or B (-50 %) or even A (-75%) compared with the average European performance level, just as is currently the case in the US Energy Star for buildings scheme. In absence of CEN standardisation on many subjects, existing methods in some countries could be provisionally approved as ways to reach an A or a B. This and others can make it possible to go quickly for the third generation suggested here. As any other standard this one would be applicable only voluntarily, here by the will of the households, companies, local authorities wishing to have buildings consuming ¾ or ½ of European average of the new buildings in 2000 and to have this environmental effort certified from outside. The satisfaction of the requirements would need to be determined at the design/installation stage through verification using simulation tools. This implies that public domain simulation tools would need to be developed to support the model building code. There are many such efforts under way in individual Member States and there would be considerable value in co-ordinating national efforts within an EU umbrella project.

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Specific recommendations The European Commission and/or a coalition of willing Member States should consider:

• the development of an EU model building code that addresses air conditioning amongst other energy end-uses. (an EU equivalent to ASHRAE 90.1 and which like ASHRAE 90.1 is subject to continuous maintenance)

• The development of practical public domain CAC system design tools which: a) can aid system designers to develop energy efficient CAC designs, b) can enable the comparison of the relative benefits of different system designs, c) can be used in building thermal regulations to demonstrate compliance with requirements

• The development of EU benchmarks for CAC system efficiency expressed in terms of: building function and size; occupancy and purpose; quality of comfort provision and climate (e.g. cooling and heating degree days)

Member States should undertake a revision of their building thermal regulations to address the following specific issues aimed at reducing CAC energy consumption:

For air-distribution systems introduce building code measures which encourage:

• Operation in mixed-mode with natural ventilation (e.g. ensuring that if ‘passive’ free-cooling is enabled mechanical cooling does not operate in those zones using free-cooling)

• The enablement of automatic free-cooling (e.g. the integration of airside and waterside economisers which are capable of operating in conjunction with mechanical cooling). Note : provisions must be included to ensure their proper functionality otherwise energy losses could occur through this measure (an obligation to do this could produce energy savings worth 5% of all CAC energy consumption by 2020: see Scenario 4 in Chapter 7)

• Efficient means being able to control air flow rates e.g. variable speed drives or variable pitch fan blades

• Proper sizing of components such as fans (e.g. requirements for maximum installed fan power (expressed as W/litre/second))

• Variable flow control (an obligation to do this could produce energy savings worth 10% of all CAC energy consumption by 2020: see Scenario 5 in Chapter 9)

• Limits on the maximum SPF of mechanical ventilation systems in new buildings (e.g. that the SPF should not exceed 1.5)

• Limits on the maximum SPF of mechanical ventilation systems in existing buildings (e.g. that the SPF should not exceed 3.0)

• Adequate sealing and insulation of ducting

• The usage of energy (heat) recovery systems

For HVAC control systems introduce building code measures which encourage:

• Restrictions on dead-bands

• Avoidance of set-point overlaps (e.g. simultaneous heating & cooling)

• Stipulations for off-hour controls including:

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1. Shutoff damper controls that automatically close when the systems or spaces served are not in use (these dampers should also satisfy a maximum allowable leakage rate.)

2. Zone isolation capabilities that permit areas of the building to continue operating while others are shut down

3. Automatic shutdown

4. Setback controls

5. Optimum start controls after the system airflow exceeds a minimum level

For refrigeration plant systems introduce building code measures which encourage:

• Free cooling from cooling towers

• Variation of fresh air using economy cycle or mixed mode

• Controls which restrict the hours of operation of the system

• Controls which prevent simultaneous heating and cooling in the same zone

• Efficient control of plant capacity, including modular plant (i.e. good part-load efficiency) (e.g. the use of power stages to allow output to be adapted to the demand)

• Efficient control of heat rejection equipment capacity, including modular plant (e.g. good part-load efficiency for cooling towers)

• Full cold thermal storage (i.e. chillers would only operate at night)

• Proper sizing of components such as pumps and refrigeration equipment

• Adequate insulation of piping

• The use of energy (heat) recovery systems

• The use of variable flow hydronic systems and wherein pumps draw substantially less power at part-load than full-load

• Heat recovery for service water heating

For CAC integrated heating systems introduce building code measures which encourage:

• Limits on the Joule heating (e.g. electric heating power provided by the Joule effect should not exceed 5% of the total heating power installed, nor 25kW by independent zone).

• Limits on terminal re-heating. (e.g. terminal re-heating is allowed for cooling-only systems but can not exceed 10% of the installed cooling power).

For the installation and commissioning of CAC systems introduce building code measures which encourage:

• commissioning tests to be conducted for boilers, chillers (power and efficiency), cooling towers, pumps, hydraulic tests, heat exchangers, controllers, noise levels and overall functionality

In addition the following specific measures which impact of CAC system design and installation are required under the Energy Performance in Buildings Directive:

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• Energy certification of new and existing buildings to verify their compliance with minimum energy performance requirements

• Certificates are to be made available to the owner or prospective tenant when the building is constructed and when it is sold or rented out. Certificates shall not be valid for more than 10 years.

• The certificate shall include reference values such as legal standards and benchmarks. It shall be accompanied by recommendations for the cost-effective improvement of the energy performance .

Recommendations about these requirements are that Member States should consider:

• Ensuring that a company or entity independent from the designer and installer should conduct the building certification

• That certifiers are trained and clear certification procedures have been developed and adopted

• That the certification would be automatically triggered for all new installations and would verify that both the installed systems and their individual components meet the energy performance requirements and attain their pre-declared performance ratings (this would require be able to measure the electrical energy, flow rates and temperatures of installed systems and their components)

• That individual items of equipment which are performing at lower than rated efficiency levels are clearly reported in the certification procedure.

9.4 Policies and measures to improve system maintenance and operation

Policy aims and potential measures targeting improved O&M The maintenance or improvement of performance, by technical measures or contractual means (Energy Performance Contracting) or by periodic audits can result in significant energy savings for CAC systems.

Optimal operation of the system can be encouraged through intelligent control regimes which in turn can be encouraged through appropriate energy performance contracting. Most end-users are unlikely to have the required expertise in house to optimise the efficient use of the CAC system and therefore it would be beneficial wthat they be encouraged to undertake suitable service contracts with specialist system operators. In fact this is already common practice today although there can be a bewildering array of contractual arrangements on offer and little access to independent assessment of the results produced. It would therefore be appropriate that policy measures be developed to encourage good practice in this regard.

It has been estimated that without proper maintenance the efficiency of CAC systems deteriorates by 2% every 5 years. As maintenance is relatively unexpensive and straightforward it is appropriate to implement policy measures which encourage regular competent maintenance to minimise this deterioration in performance. In particular the service and cleaning of heat exchangers should be properly encouraged.

Legal basis for policy measures targeting O&M Article 9 of the recent Energy Performance in Buildings Directive specifies that Member States must introduce mandatory regular inspection of AC systems above 12 kW in capacity. The inspection shall include an assessment of the AC system efficiency and sizing relative to the cooling demand of the building and advice will be provided to the users regarding possible improvement or replacement of the AC system and on alternative solutions.

Member States are obliged to ensure that this inspection is conducted in an independent manner by qualified and or accredited experts.

If implemented in a intelligent manner these requirements could go some way to ensure that existing maintenance and operation contracts are appropriate and are properly observed.

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This requirement for “independent” inspectors presumably precludes that the inspection should be conducted by the entity holding the operation and maintenance contract and therefore should provide some measure of independent verification of the proper conduct of those contracts. However, it is important that Member States implement this measure in such a way that a review of operation and maintenance contracts are encompassed within the inspection. In parallel it would be very useful that efforts be made to define best practice in operation and maintenance so that best practice guidelines can be issued against which ever existing contracts would be compared. Some aspects of these guidelines would necessarily be specific to the situation applying in individual Member States, but some would be common to all Member States. Therefore it would be very beneficial were the European Commission to take a lead in organising an EU working party charged with defining best practice in the operation and maintenance of AC systems, so that the findings could be fed into the national provisions being drawn-up by Member States. If the Commission is unable to initiate this process actors at the Member State level charged with implementing the Energy Performance of Buildings Directive could take the initiative to establish a working party of willing Member States.

The objective of the proposed Directive on Energy Demand Management (also known as Energy services) is to complete the internal market for energy by developing and encouraging energy efficiency on the demand side, especially as it is provided by utilities and service companies in the form of energy services. It is envisaged that Member States will set targets to promote and support energy efficiency services, (e.g., third party financing) and programmes, especially for smaller energy consumers such as SMEs. This could certainly be used as an opportunity to improve the O&M of CAC if appropriate rules can be defined.

Broadening the application of existing policy measures addressing O&M The Portuguese regulations impose the adoption of a maintenance plan and a monitoring system for CAC systems such that the energy consumption of all equipment with an electric power greater than 12.5 kW should be independently metered.

The UK regulations also require that the owner and/or tenant of the building be provided with a logbook that contains, amongst other things, the design assessment for CPR or other benchmarks, commissioning details, operating instructions, and details of all meters provided. There is an additional requirement that sub-metering should be provided. This includes separate metering for tenancies of more than 500 m2 (though for tenancies below 2500 m2, proportioning of cooling may be acceptable). Generally, any chiller installation (which may include more than one chiller) of greater than 20 kW input power should be separately metered, and any motor control centre providing power to fans and pumps of more than 10 kW input power.

The US ASHRAE 90.1 standard requires that drawings, manuals, and a narrative of system operation must be supplied to the building owner. This is a sensible provision because even if an engineer designs a great system, it's unlikely that energy savings will accrue if the operator doesn't know how the system should work. The standard also addresses balancing for air systems larger than 1 hp and for hydronic systems larger than 10 hp. It also requires control elements to be calibrated, adjusted, and in proper working condition for buildings that exceed 50,000 sq ft. Specific recommendations The European Commission and/or a coalition of willing Member States should consider:

• Making efforts to define best practice in operation and maintenance

• Making efforts to define best practice in operation and maintenance performance contracting

With an aim of informing national building thermal regulations and the implementation of the Energy Performance in Buildings Directive.

Member States should consider enacting measures to ensure that:

• Building owners and occupiers are provided with a logbook and an adequate operation guide for the CAC system deployed

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• A regular maintenance and monitoring system be adopted for all CAC systems (e.g. impose a requirement for regular maintenance and independent metering of CAC systems above a minimum size)

• That a competent inspectorate be developed capable of carrying the provisions of the Energy Performance in Buildings Directive applying to AC systems

• One of the roles of the inspectorate required under the terms of the Energy Performance in Building Directive should be the independent review and evaluation of CAC system operation and maintenance contracts

Member States could also consider the development of low cost finance mechanisms to encourage the adoption of good practice for CAC operation and maintenance.

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Definitions and general terms used in the study Appliance category A group of appliances or equipment that have similar technical characteristics from the

perspective of their user utility.

Categorical energy label An energy label that classifies product efficiency into one of several classes. Examples include the EU’s energy labels, which rank efficiency from A to G, and Australia’s energy label, which ranks efficiency from 1 to 6 stars. Korea, Thailand, Iran, Brazil, Mexico and India have all developed categorical energy labels.

Control cycle The period between two successive starts or two successive stops of the compressor in a refrigerating system.

Defrost cycle The period between two successive starts or two successive stops of a defrost heater in a appliance with an automatic defrost system.

Design temperature The temperature within a conditioned space that needs to be achieved during a test for the energy-consumption measurement.

MEPS Minimum energy performance standards (sometimes known as ‘minimum energy efficiency standards’).

Montreal protocol The internationally binding agreement to phase out ozone-depleting substances such as CFCs.

Net present value (NPV) The monetised value of future costs expressed in terms of their discounted value at the present time.

Payback period (PBP) The period of time it takes for a consumer to recover the extra investment made in a higher-efficiency appliance through savings in operating costs. The payback period can be ‘simple’ in that no discounting of future savings is applied, or it can be the converse in which the future savings are discounted using a real discount rate.

Thermal bridge A high thermal conductivity pathway.

Top Runner The term applied to the Japanese appliance energy-efficiency policy, wherein MEPS have been set at efficiency levels equivalent to those of the highest efficiency appliance on the market.

List of abbreviations AC Air Conditioning

ACEA EU association of car makers

ACMV Air Conditioning & Mechanical Ventilation (UK regulations)

ADENE Portuguese energy-conservation agency

AHAM US Association of Home Appliance Manufacturers

AHU Air Handling Unit

AICIA Association conducting research under the auspices of ETSIIS in Seville

AICARR Italian Association of Air Conditioning, Heating and Refrigerating Engineers.

ALCC Annualised Life Cycle Cost

ANSI American National Standards Institute

ARI American Refrigeration Institute

AS/NZS Joint test standards issued by Standards Australia and Standards New Zealand

ASHRAE American Society of Heating, Refrigerating and Air Conditioning Engineers

BAT Best Available Technology

BAU Business As Usual

BRE Building Research Establishment

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CAC Central Air Conditioners

CAHORE Cafes, Hotels, Restaurants

CEC Central and Eastern European Countries or Commission of the European Communities

CECED European major household appliance manufacturers’ association

CECOMAF See Eurovent

CEEC Central and Eastern European Countrie

CEN Committee European de Normalisation (European Committee for Standardisation)

CFC Cloroflurocarbons

COP Coefficient of Performance

CPR Carbon Performance Rating (UK regulation)

CPR Carbon Performance Rating (UK regulation)

DD Degree Days EDF Electricité de France

EER Energy-Efficiency Ratio (W/W)

EMPE Italian Method for part load rating

EMS Energy Management Systems

EPB Energy Performance in Buildings (EU regulation)

CPR Carbon Performance Rating (UK regulation)

ESCO Energy Service Companies

EU European Union

Eurovent European association of refrigeration, air-conditioning and ventilation equipment manufacturers

FCU Fan Coil Unit

GB Great Britain (excludes Northern Ireland)

GEA Group for Efficient Appliances

CPR Carbon Performance Rating (UK regulation)

HFC Hydrofluorocarbons

IDAE Spanish energy-conservation agency

IEC International Electrotechnical Committee

IPLV Integrated Part-Load Value

ISO International Standards Organisation

LBL-MIM Lawrence Berkeley Laboratory – Manufacturer Impacts Model

LCC Life-Cycle Cost

LLCC Least Life-Cycle Cost

MEPS Minimum Energy Performance Standards (same as MEES, Minimum Energy Efficiency Standards)

NPV Net Present Value

ODP Ozone-Depletion Potential

RAC Room Air Conditioners (in the wide sense)

R&D Research and Development

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RSECE The Portuguese building thermal regulations which include requirements for AC systems

SEER Seasonal Energy-Efficiency Ratio (W/W)

SPF Specific Fan Power, in Watts (of motor rating) per litre/second of airflow.

SSEER system seasonal energy-efficiency ratio (W/W)

TAC Total air conditioning (one level of comfort)

TC Total cooling (one level of comfort)

TEWI Total Equivalent Warming Impact

UK United Kingdom (includes Northern Ireland)

UoA Univeristy of Athens

US DOE US Department of Energy

US EPA US Environmental Protection Agency

VA Voluntary Agreements

VAV Variable Air Volume

VRF Variable Refrigerant Volume

VVT Variable Volume and Temperature

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AICARR,2001 Average weighed efficiency of compression chillers: AICARR’s proposal for a calculation method

AICARR, 2001, E. Bacigalupo, C Vecchio, M. Vio, M. Vizzotto., 2001, Average weighed efficiency of compression chillers: a proposal to AICARR for a calculation method, Permanent Technical Committee for "Refrigeration" in AICARR's Technical Activity Commission.

ARI, Standard 550/590, Water Chilling Packages using the vapor compression cycle, 1998.

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CEN, 1998, « Air conditioners and heat pumps with electrically driven compressors – Cooling mode”, standardCenelec EN 12055, 1998.

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CEN, 2002, prEN 14511 – 1 à 5, proposed revision of [CEN98] and [CEN97], 2002. CIBSE, 2004, "Guidance for the use of the carbon emissions calculation method: CIBSE TM32:2003" (ISBN 1 903287 41 3) Ph Davy de Virville, 1994, Control Guide for Ventilation and Air-conditioning, thermic school, Paris, France.

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REHVA, 1997 , Preliminary European Guide on HVAC design, Dominique Marchio, Eric Auzenet, ECEEE 1997 summer study proceedings

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Tiax (2002) ‘Engineering-analysis cost-efficiency curves. Commercial unitary air-cooled air-conditioners and air-source heat pumps’, TIAX LLC, Cambridge Massachusetts for the US Department of Energy, January 2002. TRIBU, (1994). "Etude Comparative des Réglementations Thermiques Européennes des Bâtiments non résidentiels", TRIBU, 1994 for Ademe

UNI, (2002) UNI Standard 10963 : Air conditioners , chillers and heat pumps- part load tests.

Westphalen, 1999) "Energy Consumption Characteristics of Commercial Building HVAC Systems Volume II: Thermal Distribution, Auxiliary Equipment, and Ventilation" Prepared by Detlef Westphalen and Scott Koszalinski, Arthur D. Little, Inc. For U.S. Department of Energy, October 1999