Analysis of the impact of adjustments to the on board HVAC design by time domain simulations

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Analysis of the impact of adjustments to the on board

HVAC design by time domain simulations

Master thesis

by

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Analysis of the impact of adjustments to the on board

HVAC design by time domain simulations

A thesis submitted in partial fulfilment of the requirements for the Master of Science degree in Mechanical Engineering in the Marine & Transport Technology department (Mechanical Systems and Integration specialisation) in the 3ME faculty at the Delft University of Technology

Conducted at AMELS BV, Vlissingen

CONFIDENTIAL

Student: Joppe G. A. Osté Student number 1380931 joppeo@gmail.com University:

Delft University of Technology

Faculty of Mechanical, Maritime and Materials Engineering (3mE)

Thesis Committee:

DELFT UNIVERSITY OF TECHNOLOGY Prof. Dr. Ir. J.J. Hopman Ir. P. de Vos

Dr. L.C.M. Itard

AMELS BV H. Koning

Ir. G.J.A. Pijnen

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ABSTRACT

The system for heating, ventilation and air conditioning (HVAC) on board of passenger related ships is the single biggest energy consumer. The system is vital to the comfort of the crew, passengers and can improve reliability of equipment. Since fuel prices keep rising and green credentials of a ship become more important it is necessary to have insight into the efficiency of the HVAC system. When a ship designer wants to analyse the potential of energy efficient measures, other HVAC layouts or needs to evaluate how the current system will behave in off-design conditions the behaviour of the system in actual service conditions is required. Power demand varies in time due to time dependent variables like ambient conditions and internal heat loads. These aspects can be investigated through time consuming on-site measuring, but this project investigates the possibility to generate this research data by time domain

simulations with an HVAC system simulation computer model. The purpose of this project is to develop a tool which provides the ship designer easy access to data of energy consumption and HVAC system behaviour in different service conditions and operational areas, in order to analyse energy efficiency in the early design stage and to gain system understanding to enable the designer to act as smart customer towards HVAC experts and suppliers.

The HVAC computer model is constructed from separate component models, each describing a physical discernible part of the HVAC process. These components are verified during extensive verification procedures, to make sure the model behaves as expected. The HVAC system of the AMELS 272 superyacht, a central variable air volume (VAV) configuration with reheating, is used as a benchmark for this project. A test case consisting of one air handling unit (AHU) and two conditioned rooms with dedicated local control is built from the component models. The benchmark simulation showed that the model is able to generate data of the energy consumption and the behaviour of the HVAC system and the separate components within a matter of seconds. From this data the energy consumption of the reheaters proved to be substantial during a one day cycle for every climate. The nominal local flow rate is determined based on the worst case design conditions and because the minimum flow rate of local supply air is limited at a certain percentage of the nominal flow rate, the local supply air provides too much cooling capacity the majority of the day. Therefore a substantial amount of reheating is required.

From the system understanding gained with the results from the benchmark simulation several suggestions arose for case studies to increase the efficiency and demonstrate the capabilities of the model to investigate the impact of adjustments in the HVAC system. By raising the AHU supply temperature with one degree up to 14% energy reduction can be realized, with the consequence that less dehumidification occurs in the AHU, resulting in higher indoor relative humidity. Lowering the flow rate limit proved to reduce the energy consumption with up to 9% for decreasing the limit with 5%, with the consequence that possibly insufficient air mixing is provided to have homogeneous temperature distribution throughout the room. Increasing the recirculation ratio reduced the energy consumption with up to 3% for every 5% extra

recirculation air. Because of the high reheating demand a measure to lower the solar heat gain through glass did not result in energy savings. To quantify the potential of the case studies more research needs to be conducted to the boundaries of the consequences of these energy efficient measures. If the consequences are found to be unacceptable, more research can be performed to measures to cope with the consequences.

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PREFACE

This document is the report for the MSc. graduation project of Joppe Osté for the MSc.

specialisation Mechanical Systems and Integration at the 3mE faculty of the Delft University of Technology. I conducted this research at the pleasant Design department of AMELS BV in Vlissingen, where I was already working at that time. Changing from employee to intern was a challenge, because sometimes it was difficult to put aside the regular work to fully focus on this project. Thanks to the support and guidance of my supervisors Hans Konings and Giorgios Pijnen I eventually managed to find the right focus and produce this thesis. Another person who supported me, even during his weekend is my Vulcanus co-member Kasper de Ruyck who helped me tremendously with his inspiration, feedback and assistance with the modelling software. Thanks to the guidance from my TU Delft supervisors Peter de Vos and Professor Hans Hopman I managed to produce the project at the desired MSc. level to graduate at the TU Delft.

I would like to thank my supervisors from the TU Delft and AMELS and Kasper de Ruyck for their time, support, guidance and assistance during this process. I want to thanks my friends, family and colleagues for their moral support and last but not least I would like to thank my girlfriend for her unconditional support, motivation and patience during this period.

If you want to have more information about this research, do not hesitate to contact me on my email address: joppeo@gmail.com

January 2016,

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LIST OF FIGURES

Figure 1-1: Electric power distribution for operational profiles on board of amels le242 ... 1

Figure 1-2: The current HVAC design process ... 2

Figure 1-3: Example daily cyclic behaviour of ambient conditions (IMTECH, 2013) ... 3

Figure 1-4: Method of research ... 4

Figure 2-1: Multi zone HVAC system lay-out, Based on (Klein Woud & Stapersma, 2011) ... 6

Figure 2-2: Central HVAC with local reheaters and vav, Based on (Klein Woud & Stapersma, 2011) ... 7

Figure 2-3: Central duo-duct HVAC system, Based on (Klein Woud & Stapersma, 2011) ... 7

Figure 2-4: Individual HVAC system with fan coil units, Based on (Klein Woud & Stapersma, 2011) ... 8

Figure 2-5: the Amels 272 ... 8

Figure 2-6: Example of air handling unit configuration ... 9

Figure 2-7: Basic compression refrigeration cycle (Fridgeman, 2008) ... 10

Figure 2-8: Adiabatic mixing of two air streams (Moran & Shapiro, 2010) ... 11

Figure 2-9: Different cooling modes ... 12

Figure 2-10: Cooling device moist air ... 12

Figure 2-11: Humidification with steam injection (Moran & Shapiro, 2010) ... 13

Figure 2-12: Typical HVAC processes in psychrometric chart (Mollier, 1923) ... 14

Figure 2-13: Layout of the HVAC system on le272 ... 16

Figure 2-14: Flow diagram temperature control with VAV/reheatING ... 17

Figure 2-15: HVAC working principles in summer conditions (Mollier, 1923) ... 18

Figure 2-16: HVAC working principles in winter conditions (Mollier, 1923) ... 19

Figure 2-17: Room internal and external heat gains and losses ... 21

Figure 2-18: Example heat load calculation of a Guest Cabin (Heinen & Hopman, 2013) ... 23

Figure 2-19: Example air distribution & reheater of guest cabin (Heinen & Hopman, 2013) ... 24

Figure 2-20: Enthalpy difference for maximum cooling capacity – Based on (Mollier, 1923)... 25

Figure 3-1: Model hierarchy AMELS HVAC system Simulink model ... 27

Figure 3-2: HVAC system LE272 with system characteristics ... 28

Figure 3-3: Submodels for a VAV HVAC system (AMELS MODEL, 2015) ... 29

Figure 3-4: Real daily ambient conditions for different regions loaded into the model ... 30

Figure 3-5: Hourly ‘SIR’ for all surface orientations in dubai (SIR calculation, 2015)... 31

Figure 3-6: Components of the AHU model (AMELS MODEL, 2015) ... 32

Figure 3-7: Mixing component (AMELS MODEL, 2015) ... 33

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Figure 3-9: Subcomponent Cooling With Dehumidification (AMELS MODEL, 2015) ... 35

Figure 3-10: Subcomponent dry cooling/heating (AMELS MODEL, 2015) ... 36

Figure 3-11: Subcomponent Steam Humidifier (AMELS MODEL, 2015) ... 37

Figure 3-12: SubComponent Fan (AMELS MODEL, 2015) ... 38

Figure 3-13: Submodel Room (AMELS MODEL, 2015) ... 39

Figure 3-14: Simulation with reheater switched on and off during the day (AMELS MODEL, 2015) ... 42

Figure 4-1: Simulation results air conditions in Psychrometric chart, with (AMELS MODEL, 2015) ... 44

Figure 4-2: SIR vs actual Solar Heat gain room at SB, heading East & North with (AMELS MODEL, 2015) ... 45

Figure 4-3: AHU supply conditions results from simulation with (AMELS MODEL, 2015)... 45

Figure 4-4: Air conditions before and after mixing for different recirc. ratios from (AMELS MODEL, 2015) ... 46

Figure 4-5: Heating and Cooling of AHU supply air at nordic climate from (AMELS MODEL, 2015) ... 46

Figure 4-6: Actual cooling capacity switching between dehumidify/dry cooling from (AMELS MODEL, 2015) ... 47

Figure 4-7: Effect of humidifier (L) and the steam flow rate (R) from (AMELS MODEL, 2015) ... 47

Figure 4-8: Fan performance simulations results from (AMELS MODEL, 2015) ... 48

Figure 4-9: Comparison Flow diagram for Local air supply control with simulation results... 49

Figure 4-10: Local Supply conditions for two rooms from (AMELS MODEL, 2015) ... 50

Figure 4-11: INdoor conditions from room 1 & 2 rooms from (AMELS MODEL, 2015) ... 50

Figure 4-12: Heat gains in summer conditions (L - Middle East) and WINTER CONDITIONS (R - NORDIC) ... 51

Figure 5-1: Benchmark input ambient conditions ... 54

Figure 5-2: Local heat distribution for Room1 (Owners office) and Room2 (Guest Cabin) ... 54

Figure 5-3: Power demand of the AHU components ... 55

Figure 5-4: Power demand of the components and the total hvac system ... 56

Figure 5-5: Benchmark energy consumption for components and total HVAC system ... 57

Figure 5-6: Relative energy consumption per component for different climates ... 57

Figure 5-7: Impact of different vav limits on the energy consumption ... 59

Figure 5-8: Impact of different AHU supply air temperatures ON energy consumption ... 60

Figure 5-9: Effect of higher AHU supply temperatures on indoor humidity ... 61

Figure 5-10: Impact of different recirculation ratios on energy consumption ... 62

Figure 5-11: Effect of higher recirculation ratio on the AHU mixed air conditions ... 63

Figure 8-1: Aerial photo of the Amels shipyard ... 68

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LIST OF TABLES

Table 1: Design conditions HVAC system (AMELS B.V., 2014) ... 15

Table 2: Total heat transfer coefficients (ISO 7547, 2002) ... 22

Table 3: Heat Gain from lighting (ISO 7547, 2002) ... 23

Table 4: Heading selection with associated orientations ... 31

Table 5: HVAC processes possible for air conditioning ... 35

Table 6: HVAC system model Controllers (AMELS MODEL, 2015) ... 42

Table 7: Load calculations according (ISO 7547, 2002) vs Simulation results from (AMELS MODEL, 2015) ... 51

Table 8: Input data for the Benchmark simulation ... 53

Table 9: Daily energy consumption for HVAC components from simulation results ... 56

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LIST OF SYMBOLS

Symbol: Description: Unit:

𝑈 Overall heat transfer coefficient [W/m2_K]

𝑅𝑖 Thermal resistance of material 𝑖 [K/W]

𝑘𝑖 Thermal conductivity of material 𝑖 [W/mK]

ℎ𝑐 Convective heat transfer coefficient [W/m

2_K]

ℎ𝑟 Radiation heat transfer coefficient [W/m

2_K]

𝐺𝑠 Solar irradiation heat gain [W/m

2_] Δ𝑇𝑟 Excess temp. by solar radiation on surface [K]

𝑄̇𝑠 Solar heat gain [W]

𝜑 Relative Humidity [%]

𝜔 Humidity ratio, unit mass of water vapour per unit mass of dry air [kg/kg]

𝑝𝑣 Vapour pressure [Pa]

𝑝𝑠𝑎𝑡 Saturation pressure [Pa]

ℎ𝑎 Specific enthalpy of dry air [J/kg]

ℎ𝑣 Specific enthalpy of water vapour [J/kg]

ℎ𝑔 Specific enthalpy of saturated vapour [J/kg] ℎ𝑓 Specific enthalpy of saturated liquid [J/kg]

𝜙𝑟𝑒𝑐 Recirculation ratio [-]

𝑇 Temperature [K]

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LIST OF ABBREVIATIONS

Symbol: Description:

AHU Air Handling Unit

FCU Fan Coil Unit

nom nominal

min minimum

max maximum

VAV Variable Air Volume

SIR Solar Irradiance [W/m2]

ACH Air Changes per Hour [1/h]

EEM Energy Efficient Measures

ESD Energy Saving Devices

SHGC Solar Heat Gain Coefficient

CC Cloud Cover [%]

Medit Mediterranean

Carib Caribbean

MidEast Middle East

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1 INTRODUCTION

1.1 Background

Analysing the operational profiles of ships with hotel functionality as core business like superyachts, cruise ships or ferries, one of the largest energy consumers is the system for heating, ventilation and air conditioning (HVAC). Because the focus of these ships is on

passenger comfort, air conditioning and indoor air quality (IAQ) is vital. Not only for passengers and crew the HVAC system is important, it can improve the reliability of (critical) equipment and as a consequence the reliability of the entire vessel. The indoor climate on board will constantly be controlled which means that the HVAC system is constantly in operation to provide acceptable indoor air quality. Therefore the HVAC system has a substantial influence on the fuel consumption and the required electric power installed on the vessel. These factors combined make the design of on board HVAC systems a challenging topic to study.

FIGURE 1-1: ELECTRIC POWER DISTRIBUTION FOR OPERATIONAL PROFILES ON BOARD OF AMELS LE242

To illustrate the impact that an HVAC system has on the total used electric power, the load distribution of all consumers of an AMELS superyacht is shown in Figure 1-1. This graph shows four operational profiles with the percentage of the estimated electric power demand per on board consumer, according to (AMELS, 2013). Manoeuvring is the only profile where HVAC is not the largest consumer, because the use of bow and stern thruster require a short peak load from the generators. Unlike the HVAC system that is constantly working the thrusters are incidentally operated.

Besides the impact on the electric power demand the HVAC system has a physical influence on the general arrangement and other technical systems like the piping system. This is why the design of HVAC is done in the early stage of the ship’s design process. The HVAC design process (Figure 1-2) starts with the design of a system concept where a decision must be made which type of HVAC layout will be implemented (Ch.2.1.1). This is typically the responsibility of the ship designer often in cooperation with HVAC subcontractors. During phase II, usually the responsibility of the subcontractor, the heat loads are estimated. Traditionally the load calculations are done based on industry standards like ISO-7547 from the International

Organization for Standardization. With data delivered by the yard like dimensions, internal heat sources and comfort requirements, the required heating and cooling capacity is estimated based

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on the maximum heat gain (for cooling) and heat loss (for heating) and worst case weather conditions. The results from the load calculations are used for the next phases; air distribution, equipment selection and duct size calculations. Eventually the power demand of the finalized HVAC design will be used in the electric load balance for the selection of electric power generators. The maximum cooling demand will be the biggest energy consuming mode of the HVAC system. This value will have substantial influence on the selection of required on board electric power.

FIGURE 1-2: THE CURRENT HVAC DESIGN PROCESS

1.2 Problem Definition

With the forecast that the world energy consumption continues to increase with more than 1.5% per year between 2015 and 2035 (SEA Europe, 2014) and the increasing scarcity of fossil fuels, it is expected that fuel prices will increase and that fuel efficiency of ships becomes more important. Another visible trend is that the green credentials of ships become more important nowadays, either because of stricter regulations, ethical reasons or sales-driven reasons.

Therefore it is necessary to have insight into the efficiency of the on board energy systems to be able to consider this aspect during the design process.

As stated in Ch.1.1 on board of a yacht the HVAC system is one of the largest electric energy consumers, which makes it an important system to analyse when research is conducted to potential efficiency improvements. The efficiency of the HVAC system can be increased with several energy efficient measures (EEMs). Possible EEMs are certain control strategies, the implementation of energy saving devices (ESD) in the HVAC system (like a heat wheel) and heat load reductions like the use of a different glass type or hull colour. Often the energy potential of ESDs are quoted in percentages, but to quantify the absolute efficiency gain of ESDs and other EEMs accurately the energy consumption in actual service conditions is required. Since the HVAC design is done in the early design stage of a ship only energy data from the load

calculations based on the most extreme conditions, used for the equipment sizing is available. For actual service conditions the transient behaviour of the system as a consequence of the thermodynamic processes, the cyclic ambient conditions (Figure 1-3), varying internal heat loads and changing solar and vessel orientation must be taken into account. Because of the time dependency of these factors it is a complex task to obtain accurate energy consumption data and provide more transparent system behaviour.

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FIGURE 1-3: EXAMPLE DAILY CYCLIC BEHAVIOUR OF AMBIENT CONDITIONS (IMTECH, 2013)

Because the early design stage of a ship is characterized by limited availability of time and resources, it is important that accurate data of energy consumption and system behaviour is easily accessible to analyse design variations without the need of expensive and time consuming experimental data.

This will lead to the following statement:

The designer must have easier access to data of the power demand and energy consumption of an on board HVAC system in actual operational conditions in order to analyse potential efficiency improvements on the HVAC system design and the system behaviour.

1.3 Research Objectives

The overall objective of the project is to develop a tool for yacht designers to gain a deeper understanding of the thermal behaviour and energy consumption of the on board HVAC system in a maritime environment. This should motivate the designer to optimize the efficiency of the HVAC system and support him in predicting the impact of HVAC design improvements and adjustments on system operation in actual service conditions without the need of experimental data. Rather than simply buy the HVAC system, the designer must be able to act as a ‘smart customer’ of the HVAC supplier and co-develop the system with the supplier in order to find the optimal system for each application. The method should:

- Enable quick investigation of the impact of implementation of certain EEMs on the actual energy consumption.

- Be able to underpin trade-off decisions during the early HVAC design process (Phase I, Figure 1-2)

- Enable investigation of the system behaviour in off-design conditions; how will the system respond if it is operating outside the worst case operational/weather conditions1_.

1.4 Methodology and Materials

To be able to quantify the energy consumption and analyse the behaviour of the HVAC system in different operational conditions and areas a computer model is created (Ch.3) which can

simulate an HVAC system in these different service conditions in a time domain, in order that

1_{In this thesis the worst case operational and weather conditions are defined as the ‘design conditions’. Although for}

most systems the design condition represent the condition in which the system operates the majority of the time, for the HVAC system it defines the extreme conditions used to calculate the required capacities (Ch.2.5)

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time dependent variables like the ambient conditions can be included. The HVAC system is divided into separate components, each describing a physical discernible part of the HVAC process. By means of simplified mathematical representations of the processes the components are created with the numerical computing environment MATLAB (R2014b) and the additional Simulink package, which adds graphical time domain simulation for dynamic systems. After dedicated verification procedures (Ch.4) the modelled components can be used to create a working HVAC system which generates the desired energy consumption data and output that provides more insight into the HVAC system behaviour.

To demonstrate the capabilities of the computer model an HVAC system test case is set up, partly based on the AMELS 272, which functions as benchmark (Ch.5.1). For a number of case studies the input parameters of the benchmark are varied (Ch.5.2) to demonstrate the

possibility to analyse the impact of potential energy efficient measures (EEMs) on the energy consumption. The complete method of research is shown in Figure 1-4.

FIGURE 1-4: METHOD OF RESEARCH

Microsoft Office Excel 2013 will be used to collect and analyse data generated from the computer model and create graphs and charts for the report, besides the plots generated by MATLAB/Simulink.

Microsoft Office Word 2013 will be used to write this thesis and other documentation during the project.

Mollier 0.4 from (Klima Delft & Van Paassen, 2000) is used for quick verification of the hygrothermal properties of moist air.

1.5 Project Limitations

This research project has a limited duration, therefore the project has several limitations: - The model components are simplified representations of HVAC processes, based on

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The level of detail of the model components is determined based on the desired result accuracy with respect to the time planning.

- The benchmark model consists of one AHU and two rooms. For future purpose it is possible to extend the model to create a total HVAC system with the components provided by this project, but this requires a method to simplify loading the yard and subcontractor input into the computer model, which is outside the scope of this project. - The variations on the benchmark are limited to four case studies:

o CASE 1: Lowering the VAV flowrate limit o CASE 2: Increasing the AHU supply temperature o CASE 3: Increasing the recirculation ratio o CASE 4: Lowering the SHGC of glass

- The validation procedure is outside the scope of this project. Since the model must be validated through on-site measurements this project is limited to verification of the computer model.

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2 HVAC SYSTEMS AND HVAC DESIGN

As an introduction to the subject and to provide more comprehensive background information about maritime HVAC systems, HVAC design and HVAC simulations this chapter will give a good overview of the data collected and reviewed literature to be able to achieve the objectives from Ch.1.3.

2.1 HVAC System

Heating, ventilation and air conditioning technology has the objective to control a set of variables in a confined space within specified constraints in such a way that human beings or objects function or stay in an optimal condition. These variables usually are: indoor air temperature and humidity, wall temperatures, air speed, carbon dioxide (CO2) concentration, purity of the air, and additionally the noise and ionization level (Van Paassen, 2004).

The ventilation system supplies air to the rooms2_{, where it must be mixed with the existing air} to reach an air condition equilibrium to realize a uniform temperature distribution in the room. Besides the air distribution to the rooms the system is exhausting air from the room to the outside and circulates the indoor air of the room. Fresh air supply is required to remove excessive moisture and CO2 produced by occupants, unpleasant odours, dust, airborne bacteria and possible smoke. For certain spaces rules and regulations define a minimum required amount of fresh air per person or in air changes per hour (MCA, 2012).

FIGURE 2-1: MULTI ZONE HVAC SYSTEM LAY-OUT, BASED ON (KLEIN WOUD & STAPERSMA, 2011)

The typical air conditioning processes are heating, cooling, dehumidifying and humidifying of air, which control the hygrothermal properties of the air. This system is interconnected with the ventilation process, because this process is responsible for the final distribution of the

conditioned air.

Systems with only ventilation functionality are encountered in engine rooms, storages and cargo spaces. Complete HVAC systems are applied to accommodation and operational spaces, rooms with sensitive equipment or possible conditioned cargo spaces.

A basic HVAC system is shown in Figure 2-1 and consists of the following main components: - Air handling unit (AHU)

2_{The vast majority of this thesis is based on a yacht, therefore spaces supplied by the HVAC system are referred to as}

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- Heat supply (electric or hot water heaters)

- Cool water supply (chilled water from the chiller unit)

- Ducting network for (fresh) air supply, room distribution, exhaust and recirculation

2.1.1 Different HVAC Layouts

Several different layouts of HVAC systems can be distinguished. The central system in Figure 2-1 lacks the possibility to control the temperature per room, which will inevitably lead to different room temperatures due to varying local heat loads. Other HVAC layouts are more likely to be installed on ships. These are some examples of HVAC systems more common for on board application according to (Klein Woud & Stapersma, 2011):

Central HVAC with local variable air volume (VAV) control and local: every room is able to regulate the amount of conditioned air supplied to the room, with the possibility to locally reheat this air. This layout is used as the case study for this project and will be discussed in more depth in the following chapters and shown below in (Figure 2-2).

FIGURE 2-2: CENTRAL HVAC WITH LOCAL REHEATERS AND VAV, BASED ON (KLEIN WOUD & STAPERSMA, 2011)

Central HVAC with duo-duct system (Figure 2-3): Supplies a separate hot and a cold air stream to every room that makes it possible to mix these locally to obtain required air condition. This system has the advantage that minimum automatic control is required and there is no need for local heating and cooling equipment. The major drawback of this system is the inability to finely control temperature and more importantly humidity, which is especially for passenger vessel applications reason to exclude this system (ASHRAE CH11, 2007).

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Individual AC system (Figure 2-4): every room is equipped with a small local AHU, usually referred to as fan coil unit (FCU). The fresh air supply is provided by a central fresh air unit (FAU). The major advantages of this system are good individual control and small air ducts throughout the ship. The drawback of this system is that it tends to be more expensive than the alternatives and every room needs dedicated hot and cold water supply for their FCU, resulting in more piping infrastructure.

FIGURE 2-4: INDIVIDUAL HVAC SYSTEM WITH FAN COIL UNITS, BASED ON (KLEIN WOUD & STAPERSMA, 2011)

2.1.2 Reference HVAC system AMELS 272

Since this project is done at AMELS shipyard a superyacht from the AMELS fleet is used as reference for this project. The HVAC system biggest yacht in the current AMELS fleet; the 83 meter long Limited Edition 272 (Figure 2-5) is investigated as a case study for this project. As stated above, this superyacht is equipped with central HVAC layout with hot water reheaters and VAV regulation. Four systems with a layout according to Figure 2-2, all with dedicated AHU, supply four separate zones on board to keep the ducting network of acceptable length and diameter and provide more redundancy. More information about this system can be found in APPENDIX 1 and APPENDIX 2.

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2.2 Main Components

An HVAC system is built up from the main components. They are the basis of the system, determine the design of the rest of the system and have a big influence on the system performance.

2.2.1 Air Handling Unit

The AHU is the core of an HVAC system. It conditions the air and supplies this conditioned air toe the rest of the HVAC system. Figure 2-6 shows an example of an AHU, also known as air conditioning unit (ACU) used in an HVAC system with reheating and VAV functionality (as seen in Ch.2.1.1). An AHU is traditionally equipped with:

1. Mixing section – mixing fresh and recirculated air 2. Filtering section

3. (Pre)Heating section – electric heater

4. Cooling section – cooling coil with chilled water from the chiller unit

5. Fan section – belt driven centrifugal fan with electric motor. Placed in front of humidifier. Often this is the other way around

6. Steam humidification section

7. Distribution section – plenum that has been insulated against sound and heat dissipation

Not part of the AHU but for the VAV functionality as seen in Figure 2-2 every room must room be equipped with a dedicated:

8. Reheating section – electric (or ‘wet’) heater in the beginning of the supply duct to the accommodation

9. Air volume control – flow damper in the beginning of the supply duct to the accommodation

FIGURE 2-6: EXAMPLE OF AIR HANDLING UNIT CONFIGURATION The working principles of the system from Figure 2-6 are as follows:

The AHU receives an amount of fresh and recirculated air. These flows are mixed according to the set recirculation ratio, filtered and conditioned to the required distribution condition

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(Figure 2-6.7). From there the air is being distributed to the rooms. To be able to handle variations in the air demand per room the fan is controlled to keep the air pressure constant in the distribution section. The air condition in the distribution section is set to be kept constant (e.g. on the AMELS 272 this is set to approx. 12°C). In this specific configuration the local indoor temperature is controlled by regulating the air flow rate (Figure 2-6.9) with the possibility to reheat the air (Figure 2-6.8) when the room temperature is getting too low and minimum air supply is reached. More about supply air flow rate constrains in Ch.2.4.5.

2.2.2 Heat Supply

There are two different methods of heat supply for on board HVAC systems; global or local heat supply. Global heat supply is the heat fed to the AHU to heat up the air stream going through the AHU. The local heat supply refers to the dedicated heat supplied in the supply duct to a single space to be able to control each space separate. This is not applicable to HVAC system layouts like the Duo Duct system in Figure 2-3, but for most of the on board HVAC systems local heat supply is applied.

For both global and local two types of heating are possible: - Electric heating

- Hot water heating, supplied by boilers

2.2.3 Chilled Water Supply

In the AHU the cooling energy is supplied by means of chilled water flowing through cooling coils placed in the air stream. This chilled water is supplied by the chilled water plant, which is produced according to the compression refrigeration process. A simple schematic diagram of this process is shown in Figure 2-7.

The performance of this cooling process is defined as the coefficient of performance (COP). This COP is the ratio between the heat received by the chilled water and the required electric power demand of the compressor:

𝐶𝑂𝑃𝑐ℎ𝑖𝑙𝑙𝑒𝑟 =

𝑄̇𝑐ℎ𝑖𝑙𝑙𝑒𝑑𝑤𝑎𝑡𝑒𝑟 𝑃𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟

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2.3 HVAC Processes

To be able to study HVAC design and HVAC system behaviour it is important to know the mathematical approximations of the physical processes that occur in HVAC systems. In this section these approximations are reviewed. The most important properties and behaviour of moist air can also be analysed by a graphical representation that is provided by psychrometric charts. In Figure 2-12 the typical HVAC processes are drawn into the psychrometric chart to make it easier to analyse the processes. By connecting the lines for occurring processes a whole system can be drawn into the graph to determine the physical and thermodynamic properties of the moist air during the entire process.

During this project assumptions have been made to simplify the calculations, as it is known from thermodynamic handbooks that the following effects are of minor importance for the results (Moran & Shapiro, 2010):

- Effects of kinetic and potential energy will be ignored

- Total pressure remains constant at 1.013 bar throughout the HVAC processes - Inertia of air will be ignored

2.3.1 Adiabatic Mixing of Two Air Streams

The first common process in the HVAC system is the mixing of the ambient air stream with the recirculated air stream. These two air streams are mixed to a single air stream with resulting thermodynamic properties (Figure 2-8). The recirculation ratio (

𝜙

𝑟𝑒𝑐) will define the amount of recirculation and fresh air that will be mixed. There is no heat, mechanical power or moist being added or extracted, therefore the process is assumed to be adiabatic and visualized in Figure 2-12 line 1. The point where the arrows meet represents the state of the mixed air.

FIGURE 2-8: ADIABATIC MIXING OF TWO AIR STREAMS (MORAN & SHAPIRO, 2010)

The mathematical representation derived from the mass and energy balance (Moran & Shapiro, 2010) is: 𝒎̇𝒂𝟏(𝒉𝒂𝟏+ 𝝎𝟏𝒉𝒈𝟏) + 𝒎̇𝒂𝟐(𝒉𝒂𝟐+ 𝝎𝟏𝒉𝒈𝟐) = 𝒎̇𝒂𝟑(𝒉𝒂𝟑+ 𝝎𝟑𝒉𝒈𝟑) Where3_: 𝒉𝒈(𝑻) ≈ 𝒉𝒗 𝑯𝒊= 𝑯𝒂,𝒊+ 𝑯𝒗,𝒊= 𝒎𝒂,𝒊𝒉𝒂,𝒊+ 𝒎𝒗,𝒊𝒉𝒗,𝒊

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2.3.2 Heating or Dry Cooling

Heating is a relatively easy process, because there is no addition or extraction of moist. It is graphically represented in Figure 2-12 as an upward vertical line at constant humidity ratio ω. Cooling is a more complex process. Two forms of cooling are possible (Figure 2-9); cooling without the occurrence of condensation called dry cooling or sensible cooling, and cooling with the occurrence of condensation called dehumidification. This form is described in the next section. When exclusively sensible cooling occurs the wall of the cooler is above the dew point4 temperature of the air and no dehumidification will occur. This form of cooling is represented in Figure 2-12 as a downward vertical line, the opposite of heating. For steady-flow conditions both processes can be calculated with:

𝑄̇𝑠𝑒𝑛𝑠 = 𝑚̇𝑎𝑐𝑝,𝑎Δ𝑇 = 𝑚̇𝑎Δℎ

For cooling the temperature and enthalpy difference is negative, hence the energy flow rate is negative for cooling and positive for heating.

FIGURE 2-9: DIFFERENT COOLING MODES

2.3.3 Dehumidification

When a moist air stream is cooled at constant mixture pressure by a cooling coil to a temperature below its dew point temperature, dehumidification of the air will occur. This means that from a certain (dew point) temperature condensation of the water vapour initially present will occur and the humidity ratio ω will decrease. The heat extracted from the air for the (phase changing) condensation process is called latent cooling and will not result in a

temperature change. Line 3 in Figure 2-12 shows the graphical representation of the process that is a combination of sensible cooling (vertical part) and dehumidification (rest of the line along the saturation line).

FIGURE 2-10: COOLING DEVICE MOIST AIR

4_{The dewpoint temperature is the temperature to which air must be cooled down to reach 100% relative humidity}

assuming there is no evaporation into the air. At this temperature the air must condense water vapour into liquid water when cooled down any further

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With mass and energy balances a mathematical representation (Moran & Shapiro, 2010) for the rate of heat transfer between the moist air stream and the refrigerant coil (𝑄̇_{𝑑𝑒ℎ𝑢𝑚}) like in Figure 2-10 can be found:

𝑄̇𝑑𝑒ℎ𝑢𝑚= 𝑚̇𝑑𝑎[(ℎ𝑎2− ℎ𝑎1) − 𝜔1ℎ𝑔1+ 𝜔2ℎ𝑔2+ (𝜔1− 𝜔2)ℎ𝑓2]

The condensate (𝑚̇_𝑤) is assumed to be drained as a saturated liquid at T2, which means ℎ_𝑤= ℎ𝑓2.

2.3.4 Humidification

It is often necessary to increase the moisture content of the air in occupied spaces. Several methods of humidification exist, but for on board purposes steam humidification is used almost exclusively.

With steam-spray humidification steam with relatively high temperature is injected into the air stream. Because this is not an adiabatic process (evaporation heat is added to the air) not only the humidity ratio but also the air temperature will increase, as shown in Figure 2-12 line 4.

FIGURE 2-11: HUMIDIFICATION WITH STEAM INJECTION (MORAN & SHAPIRO, 2010)

According to (Moran & Shapiro, 2010) this process of humidification by steam injection can be described with the mass and energy balances as:

(ℎ𝑎+ 𝜔ℎ𝑔)₂= (ℎ𝑎+ 𝜔ℎ𝑔)₁+ (𝜔2− 𝜔1)ℎ𝑔3

The first term on the right can be obtained with the hygrothermal properties of the initial air (at 1). For the second term the mass flow rate of steam injected must be known to determine the humidity ratio rise (𝜔₂− 𝜔1). Then the hygrothermal properties of the final air (at 2) are known.

2.3.5 Psychrometric Charts

A graphical representation of several important properties of moist air is provided by

psychrometric charts. These charts, also known as Mollier diagram (Klein Woud & Stapersma, 2011), shows the specific enthalpy (ℎ in [kJ/kg] as a diagonal axis) as a function of the humidity ratio (𝜔 in [kg/kg] on the horizontal axis). Other properties that are provided in the charts are:

- Dry bulb temperature (𝑇) on the vertical axis

- Wet bulb and dew point temperature (𝑇_𝑤𝑏); sloped axis along 𝜙 = 100% - Relative humidity (𝜙 in %)

As described in the previous section, this psychrometric chart can be used to visualise and analyse the HVAC processes. In Figure 2-12 these processes and a comfort area is drawn. The (ISO 7730, 1984) standard defines comfort as “That condition of mind which expresses

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parameters that can be measured. Therefore different size comfort areas can be defined. To comply with the high AMELS comfort standards this area is kept small as shown in Figure 2-12.

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2.4 HVAC System AMELS 272

As a reference for this project the HVAC system on the AMELS 272 is taken. In this chapter this system that is under investigation will be described more elaborately.

In the AMELS building specifications of the LE272 (AMELS B.V., 2014) the design criteria for the HVAC system are defined by the shipyard:

Summer conditions: Winter conditions:

Outside Temperature 35°C 0°C

Outside Relative Humidity 75% 80%

Inside Temperature 22°C 22°C

Inside Relative Humidity 55% (max.) 35% (min.)

Seawater Temperature 32°C 4°C

Recirculation Air Amount 50% 50%

Min. Fresh Air per Person 30 m3_/h _{30 m}3_/h

Air Changes per hour (ACH) 9 9

TABLE 1: DESIGN CONDITIONS HVAC SYSTEM (AMELS B.V., 2014)

With the design criteria from Table 1 the yard will choose in cooperation with specialists and subcontractors an HVAC configuration based on several design considerations like practicality, flexibility, efficiency, size, piping footprint and budget constraints. As stated in Ch.2.1.1 the AMELS 272 is equipped with multiple central HVAC systems with local reheaters and VAV regulation. In APPENDIX 1 an overview of one of these systems is shown and APPENDIX 2 shows how the system eventually is installed on board in the 3D model of the AMELS 272. In the next section the HVAC systems will be described in more detail.

2.4.1 System Characteristics

The air conditioning system characteristics on board of the LE272 are of a single duct reheat type with variable air volume (VAV) distribution (as shown in Figure 2-13). For the VAV system the electric motors of the fans are speed controlled through frequency inverters to maintain constant pressure in the plenum (between the fan and the VAV control). Every room is equipped with dedicated VAV control and reheater, which are centralized close to the AHU usually in a fan room. From the fan room the ducting network distributes the air to the local diffusers or grilles to supply the air to the rooms. The system is automatically controlled, with the extract system connected to avoid under pressure in the accommodations. Overpressure from accommodation shall direct recirculation air via overpressure ducts to the passages back to the AHUs and the exhaust air via sanitary spaces. It is possible to create a “night modus” in which AHUs are switched to 50% capacity as well as the extract system. The humidity of the indoor air shall be controlled in the AHUs by steam humidifiers.

Some spaces like pantries, the galley, technical spaces (like steering gear compartment) shall be provided with FCU’s for various reasons. They will cool the air by means of cooling coils fed by chilled water and heat the air by means of an electric heater. Each FCU is provided with a control unit to adjust room temperature and fan speed. The fresh air will be delivered from a fresh air unit (FAU) instead of an AHU.

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FIGURE 2-13: LAYOUT OF THE HVAC SYSTEM ON LE272

2.4.2 Working Principle Summer Conditions

The working principles of the HVAC system in summer conditions are as follows:  The chilled water (CW) plant will be in operation

 The CW pumps will deliver chilled water of 6°C to the AHUs to cool the mixture of fresh and recirculated air down to the AHU supply temperature: 𝑇_𝐴𝐻𝑈= 12°𝐶

 The cooled AHU supply air will be distributed by the single duct system to the local diffusers or grilles in the spaces where by means of electrically controlled air volume flow dampers (VAV valves) the air volume flow rate and thus the room temperature will be controlled as shown in Figure 2-14. The control method for this system

(VAV/reheater configuration) works as follows:

o When room temperature drops below indoor temperature set point (𝑇_𝑠𝑒𝑡= 22°𝐶) the VAV valve gradually closes to decrease the air flow (reheating is off and 𝑇_{𝑠𝑢𝑝𝑝𝑙𝑦}= 𝑇𝐴𝐻𝑈 = 12°𝐶). The VAV valves are limited to a minimum flow rate of 70% of the nominal flow rate (more information about this limit in Ch.2.4.5). o When room temperature is still decreasing at minimum VAV flow rate (70% of

nom) the reheater will heat the AHU supply air (𝑇_𝐴𝐻𝑈= 12°𝐶) to the required local supply temperature so that 𝑇_{𝑠𝑢𝑝𝑝𝑙𝑦}> 12°𝐶, until 𝑇𝑠𝑒𝑡 is reached.

o When the room temperature exceeds 𝑇𝑠𝑒𝑡 reheating is gradually turned off. When the reheater is fully turned off the VAV valve will take over temperature control and gradually open until 𝑇_𝑠𝑒𝑡 is reached with 𝑇_{𝑠𝑢𝑝𝑝𝑙𝑦}= 𝑇𝐴𝐻𝑈= 12°𝐶 This is a continuous process which means that the VAV valves are constantly moving to control the room temperature with AHU supply air (𝑇_𝐴𝐻𝑈= 12°𝐶) unless reheating is necessary. In that case the VAV valves are at minimum VAV flow rate position (70% of nom) and reheating control will take over.

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FIGURE 2-14: FLOW DIAGRAM TEMPERATURE CONTROL WITH VAV/REHEATING

2.4.3 Working Principle Winter Conditions

The working principles in winter conditions are basically the same, however at extremely low temperatures:

 The CW plant will not be in operation

 The mixture of fresh and recirculating air will be fully automatically controlled heated up to 𝑇_𝐴𝐻𝑈= 12°𝐶 by the preheaters and potentially (steam) humidified in the AHU to desired supply humidity. The temperature control after that is as described in the last step of 2.4.2 and shown in Figure 2-14; in practice this means that the flow dampers are at minimum flow rate with the reheaters in operation, reheating the AHU supply air of 12°C to temperatures of max. 35°C (in Figure 2-16 stations5_{7) before supplied to the}

room.

2.4.4 Working Principles in Psychrometric Charts

From Ch.2.3.5 it is known that the HVAC processes can be visualized in the Psychrometric chart for moist air. To elaborate the working principles of the reference HVAC system as described in the previous sections both principles (in summer and winter conditions) are plotted in Figure 2-15 and Figure 2-16. The numbers in both charts refer to in the HVAC process in Figure 2-13. Note that between supply condition (station 7) and indoor condition (station 8) there is a slight increase in the humidity ratio because of the moist production from humans in the room.

5_{A station is an air condition at some point in the HVAC process which can but does not exclusively have a different}

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FIGURE 2-16: HVAC WORKING PRINCIPLES IN WINTER CONDITIONS (MOLLIER, 1923)

2.4.5 Control and Flow Dampers

The temperature measured in the room (usually in the overpressure ducts) will be fed back to the local controller. The controller will maintain the indoor temperature set point by operating the electrically controlled VAV valves and the reheaters in the supply duct.

As mentioned before, the valve position is limited to a minimum air flow rate. This limitation makes sure there is proper air mixing throughout the room. Especially in winter conditions, when hot air is supplied to the room, a certain air speed is required for a certain flow pattern

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which makes sure the warm air reaches the lower part of the room to get a uniform

temperature distribution throughout the room. In the case that electrical reheaters are used overheating of the heaters can occur when not enough heat is extracted by a low air flow rate. The AHU supply air temperature of 12°C in the distribution section after the AHU will be automatically controlled by the AHU as described in Ch.2.2.1 and is valid for both summer and winter conditions.

2.5 HVAC System Design Method

The current AMELS method of the HVAC system design, with the focus on the load calculations (Phase II in Figure 1-2), is described in this section for the AMELS 272. For this yacht the load calculations by (Heinen & Hopman, 2013) are based on (ISO 7547, 2002). The ISO standard 7547 specifies design conditions and methods of calculation for marine HVAC systems for all conditions except those encountered in extremely cold or hot climates (i.e. within the AMELS design criteria from (AMELS B.V., 2014)).

The results that follow from the ISO standard are: - The required air volume flow rates - The required air changes per hour - The internal and external heat loads

- The local and total required cooling and heating capacity

2.5.1 Load Calculations

To be able to calculate the required cooling and heating capacity the internal and external heat gains and losses per room, shown in Figure 2-17 must be determined (ASHRAE CH11, 2007). The heat loads that are estimated are:

- Solar radiation

- Heat transmission through hull, decks and bulkheads - Heat (latent and sensible) dissipation from occupants - Heat gain from lights

- Heat gain from equipment

The following input is required to calculate the heat loads:

A. Physical dimensions; ceiling height and surface areas of bulkheads (either to adjacent rooms or to the outside), floor, ceiling and windows

B. Colour (light/dark) and orientation (vertical/horizontal) of outdoor facing surfaces C. Maximum occupancy of the room

D. Other heat sources (like AV, IT or bridge equipment), usually defined as ‘apparatus’ The methods of calculation for the factors discussed above based on (ISO 7547, 2002) are described in the following sections.

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FIGURE 2-17: ROOM INTERNAL AND EXTERNAL HEAT GAINS AND LOSSES

2.5.1.1 Heat Transmission through Hull, Decks and Bulkheads

The heat transmission between rooms and to the outside is determined with the physical dimensions, the total heat transfer coefficients (𝑈) defined by ISO.7547 for different surfaces (Table 2) and the temperature difference. In (Heinen & Hopman Eng. BV, 2013) the temperature difference (Δ𝑇) is taken as:

- Between adjecent spaces: Δ𝑇 = 3𝐾

- Outside summer conditions: Δ𝑇 = 𝑇𝑜𝑢𝑡− 𝑇𝑖𝑛 = 35 − 22 = 13𝐾 (heat gain) - Outside winter conditions: Δ𝑇 = 𝑇𝑜𝑢𝑡− 𝑇𝑖𝑛 = 0 − 22 = −22𝐾 (heat loss) The heat transmission between spaces and the outside can be determined with:

𝑸̇𝒕𝒓= 𝑼 × 𝑨 × 𝚫𝑻

Because the worst case scenarios are estimated (eventually the max. required heating and cooling capacity is determined) these figures must have negative influence on the scenario that is being estimated. This means that in winter conditions heat transmission is heat loss to outside and adjecent spaces, and in summer conditions heat transmission is heat gain from outside and adjecent spaces.

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TABLE 2: TOTAL HEAT TRANSFER COEFFICIENTS (ISO 7547, 2002)

With the sum of these heat losses by means of transmission for all surfaces the required heating capacity for winter conditions is found, because this are the only heat losses from a room. For the required cooling capacity more heat sources must be taken into account:

- Solar radiation

- Heat gain from persons; occupancy - Lighting

- Apparatus

2.5.1.2 Solar Heat Gain

For outer bulkheads (Heinen & Hopman, 2013) uses the equation for heat transmission from Ch.2.5.1.1 with the total heat transfer coefficient from Table 2, but with a Δ𝑇 for the excess temperature (above the outside temperature of 35°C) caused by the solar radiation on surfaces with data from 2.5.1B:

o For vertical light surfaces: Δ𝑇𝑟 = 12 [𝐾] o For vertical dark surfaces: Δ𝑇𝑟 = 29 [𝐾] o For horizontal light surfaces: Δ𝑇_𝑟 = 16 [𝐾] o For horizontal dark surfaces: Δ𝑇_𝑟 = 32 [𝐾] Solar heat gain through glass surfaces is determined with:

𝑄̇𝑠𝑜𝑙,𝑔𝑙𝑎𝑠𝑠 = 𝐴 × 𝐺𝑠 Where 𝐺𝑠 is the heat gain per square meter from glass surfaces:

o For clear glass surfaces: 𝐺𝑠 = 350 [𝑊/𝑚2]

o For clear glass surfaces with interior shading: 𝐺𝑠 = 240 [𝑊/𝑚2]

2.5.1.3 Occupancy, Lighting and Apparatus

The total heat gain from occupants is determined by multiplying the number of persons in the room with the average sensible heat emitted by a person at rest; 70W.

For the heat gain from lighting ISO.7547 uses standard values of average heat emitted by lighting per square meter of floor. Rooms with different functions can have different values for heat gain from lighting and a distinction has been made between incandescent and fluorescent light as shown in Table 3.

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TABLE 3: HEAT GAIN FROM LIGHTING (ISO 7547, 2002)

Finally the heat emitted by the equipment in the room must be added, which is estimated by the shipyard. With all these heat gains the total required cooling capacity can be calculated.

Figure 2-18 shows an example of such a load calculation for a guest cabin. Clearly the total required cooling capacity is higher than the required heating capacity, because for heating only heat losses through transmission (626W in Figure 2-18) are accounted for, while for cooling also the other heat gains described above are accounted for (-1642W in Figure 2-18). With this data phase III and IV from Figure 1-2, respectively the air distribution and the equipment selection follow.

FIGURE 2-18: EXAMPLE HEAT LOAD CALCULATION OF A GUEST CABIN (HEINEN & HOPMAN, 2013)

2.5.2 Air Distribution and Reheating

When the required heating and cooling capacities are known from the method described in Ch.2.5.1 the air distribution values can be determined. The highest required volume flow, either for heating or cooling is being used for the sizing of equipment and ducts, provided that this flow rate exceeds the minimum required ACH specified by the shipyard (AMELS B.V., 2014) or rules & regulation like (MCA, 2012). If this is not the case the specified ACH is applied.

From the load calculations of the example in Figure 2-18 the required air volume flow rate for maximum heating and cooling can be determined:

𝑉̇𝐶/𝐻=

𝑃𝐶/𝐻 𝜌𝑎𝑖𝑟∙ Δ𝑇𝑚𝑎𝑥∙ 𝑐𝑝

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With a maximum temperature difference between heating or cooling Δ𝑇_𝑚𝑎𝑥= 10 [K], for a supply air temperature of 32°C and 12°C for respectively heating and cooling the room to the desired indoor temperature of 22°C.

When the required flow rates (𝑉̇_𝐶/𝐻) are divided by the total volume of the room, the ACH for maximum heating and cooling are known. The highest ACH will represent the actual required ACH and associated actual required air quantity as shown in Figure 2-19.

FIGURE 2-19: EXAMPLE AIR DISTRIBUTION & REHEATER OF GUEST CABIN (HEINEN & HOPMAN, 2013)

The design condition for the reheater (winter) takes into account: - No internal heat gains

- No solar heat gain

- Maximum heat losses (TOTAL HEATING CAPACITY REQUIRED in Figure 2-18) due to lower ambient temperature and temperatures in adjacent rooms.

To determine the required capacity for the reheater the minimum air volume flow rate (𝑉̇_{𝑟𝑒ℎ𝑒𝑎𝑡}) is required. This flow rate is fixed for the reheater in operation at minimum VAV flow, because when the flow rate will increase the reheater is already switched off (see Figure 2-14). With this fixed flow rate (70% of actual required air quantity, see Ch.2.4.5) the required reheating

capacity can be calculated. The reheater must add enough heat to the air stream to be able to compensate for the maximum heat losses from the room and the cooling capacity in the air stream from the AHU:

𝑃𝑟𝑒ℎ𝑒𝑎𝑡= 𝑄̇𝑙𝑜𝑠𝑠+ (𝜌𝑎𝑖𝑟∙ 𝑉̇𝑟𝑒ℎ𝑒𝑎𝑡∙ 𝑐𝑝,𝑎𝑖𝑟∙ (𝑇𝑖𝑛− 𝑇𝐴𝐻𝑈)) Where; 𝑉̇_{𝑟𝑒ℎ𝑒𝑎𝑡}= 70% × 𝑉̇𝑚𝑎𝑥

2.5.3 Equipment Selection

With the data collected in previous sections the equipment can be selected. The reheaters can be selected based on their required heating capacity. The fans can be selected with the known flow rates.

With the sum of the maximum required air quantity for all the rooms that are supplied by the AHU and known air supply conditions (Ch.2.4.2), the total required cooling and heating capacity can be calculated. The maximum cooling and heating capacity for the AHU is the product of the maximum total air mass flow and the enthalpy difference between intake air and supply air conditions. This enthalpy difference can be measured from the psychrometric chart (Figure 2-12) or calculated with the equations in Ch.2.3.

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To illustrate this, one of the AHU’s from the AMELS 272 is analysed.

The maximum total air volume flow rate for the AHU (sum of all the rooms): o 𝑉̇ = 7550 [m3_/hr]

The supply air condition: o 𝑇𝑠𝑢𝑝𝑝𝑙𝑦= 12°C o 𝜑𝑠𝑢𝑝𝑝𝑙𝑦≈ 99%

The air condition at AHU intake: mixture of 50% recirculation air and 50% outside air in summer:

o 𝑇_{𝑟𝑒𝑐𝑖𝑟𝑐} = 22°C & 𝑇𝑠𝑢𝑚𝑚𝑒𝑟 = 35°C  𝑇𝑖𝑛𝑡𝑎𝑘𝑒= 28.5°C

o 𝜑_{𝑟𝑒𝑐𝑖𝑟𝑐} = 50% & 𝜑𝑠𝑢𝑚𝑚𝑒𝑟 = 75%  𝜑𝑖𝑛𝑡𝑎𝑘𝑒≈ 71% (from Figure 2-12) And outside air in the winter:

o 𝑇_{𝑟𝑒𝑐𝑖𝑟𝑐} = 22°C & 𝑇𝑤𝑖𝑛𝑡𝑒𝑟= −5°C  𝑇𝑖𝑛𝑡𝑎𝑘𝑒= 8.5°C

o 𝜑_{𝑟𝑒𝑐𝑖𝑟𝑐} = 50% & 𝜑𝑤𝑖𝑛𝑡𝑒𝑟= 80%  𝜑𝑖𝑛𝑡𝑎𝑘𝑒≈ 75% (from Figure 2-12) Enthalpy differences from Figure 2-20:

o Δℎ_{𝑐𝑜𝑜𝑙} ≈ 43 [kJ/kg] o Δℎ_{ℎ𝑒𝑎𝑡}≈ 5 [kJ/kg]

This results in the required capacities for the AHU: o 𝑄̇_𝐶 = Δℎ𝑐𝑜𝑜𝑙∙ 𝑉̇𝜌𝑎𝑖𝑟∙ 1 3600= 108𝑘𝑊 o 𝑄̇_𝐻= Δℎℎ𝑒𝑎𝑡∙ 𝑉̇𝜌𝑎𝑖𝑟∙ 1 3600= 12.5𝑘𝑊

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The selection of the fans used in the AHU is based on the total air volume flow rate for the AHU and the total static pressure loss of the air flow. The static pressure loss is the sum of external (ductwork, diffusers) and internal (from AHU) pressure losses.

With the gathered data the AHU’s can be selected. With the total required cooling capacity for all the AHU’s the HVAC related CW plant capacity is known. For the CW plant selection the cooling demand from other systems, like cool and freeze stores, cooling of (electric) equipment or potable water is required to determine the total chilled water capacity.

Several assumptions in the ISO standard 7547 for the load calculations will most likely result in overcapacity of the total HVAC system, because:

- The solar irradiation can never be equal on surfaces in all directions. Inevitably there are surface areas in the shade, while according to (ISO 7547, 2002) the same solar heat gain exists on all surface areas at the same time.

- The heat transmission between adjacent rooms is calculated with Δ𝑇 = 3𝐾 between the rooms, which is impossible. In practice all adjacent rooms (exceptions excluded) are conditioned by the HVAC system and shall have comparable indoor conditions, which results in a minor temperature difference between adjacent rooms. Therefore there will be negligible heat transmission between adjacent rooms.

- Maximum heat dissipation from lighting, equipment and occupancy won’t occur in all rooms at the same time.

This is currently solved by multiplying the total required cooling capacity (𝑄̇_𝐶) by a diversity factor 𝑘𝑠 (or simultaneity factor):

𝑄̇𝐶,𝑖𝑛𝑠𝑡𝑎𝑙𝑙𝑒𝑑= 𝑘𝑠∗ 𝑄̇𝐶 Where: 𝑘_𝑠 = 0.7

The diversity factor currently applied at AMELS is an empirical value, resulting from extensive on board measurements by the subcontractor. This value is one of the interesting aspects that can be evaluated with the result of this project.

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3 MODELLING OF THE HVAC SYSTEM

In many fields, simulation models have become established tools used during the design process. For the design of on board HVAC systems a simulation model can be used to provide more insight into system behaviour, control strategies and energy consumption. With the data generated by such a model system evaluation, optimization and failure identification can be performed in the early stages of the design process. According to (Grimmelius & Zheng, 2008), when creating an HVAC simulation model it is not essential to model components like rooms, heat exchangers and chillers in detail to gain more understanding of the system behaviour, as long as the mass and energy balances are respected. A practical approach to model a complex dynamic system like an HVAC system, is to divide the system in smaller parts, each describing a physical discernible part of the process. This approach is applied to the model for this project. The model hierarchy for this project is as follows:

FIGURE 3-1: MODEL HIERARCHY AMELS HVAC SYSTEM SIMULINK MODEL

The HVAC system model is created by connecting submodels that consists of components, which can contain subcomponents, depending on the depth of the submodel. The simplified (sub) component models can be of the ‘first principle’ type or the ‘fit’ type. The majority of the models are first principle models, which are based on physics. Fit models use actual performance data of manufacturers, like fan models based on performance curves delivered by the manufacturers. For this project the HVAC system of the AMELS 272 (Figure 3-2) functions as a benchmark for the model. The following system characteristics are implemented in the model:

- The indoor air in a room is influenced by internal and external heat gains and losses (Figure 2-17) which influence the indoor air temperature and relative humidity in the room

- The indoor air temperature and humidity is controlled by the supply of heat or cooling energy to compensate for the heat gain/loss to or from the room

- The heat supply is realised by a conditioned and potentially reheated air stream, of which the amount of air is controlled by the VAV valve (working as described in Ch.2.4.2)

- The air stream to the rooms is conditioned by the AHU to a predefined air condition - The AHU will mix a predefined quantity of recirculation air (indoor air conditions) and

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FIGURE 3-2: HVAC SYSTEM LE272 WITH SYSTEM CHARACTERISTICS

In this chapter the submodels and its (sub) components are described according to the following aspects:

I. Purpose of the model II. Input signals of the model III. Output signals of the model IV. Model Parameters

3.1 Overall HVAC System Model

As stated before, the purpose of the HVAC system is to control a set of variables in the spaces on board within the design conditions in such a way that human beings feel comfortable or objects can function or stay in an optimal condition. The controlled variables that the overall HVAC system model focusses on are therefore:

- Indoor air temperature of the room(s) - Humidity in the room(s)

The HVAC system model simulates the system in the time domain to obtain more insight into system behaviour under actual operating conditions. Eventually the desired output for evaluation will be the energy consumption of the complete HVAC system for a given period of time. With these results the efficiency of the system in different operational and ambient conditions can be evaluated and potentially optimized.

To be able to simulate the complete HVAC system the required submodels are:  Model for ambient and ship conditions

 AHU model  Room model(s)

 Local supply air control model(s) per room  Exhaust ducting model

 Chiller unit

A collection of these submodels is shown in Figure 3-3 and the whole system model can be found in APPENDIX 3.

The parameters for the system model are different operational conditions like the area of operation, the heading of the vessel or on board occupancy.

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FIGURE 3-3: SUBMODELS FOR A VAV HVAC SYSTEM (AMELS MODEL, 2015)

3.2 Submodel: Ambient and Ship Conditions

The purpose of the Ambient Model is to define the ambient conditions. The input signal for the model is the current time step. The output signals of this model are:

- Outside temperature - Outside humidity - Solar radiation

And the constants for ambient conditions: - Atmospheric pressure

- CO2-concetration in the air

The outside temperature, humidity and atmospheric pressure define the ambient air conditions that functions as input for the AHU. To calculate the heat transfer between indoor and outdoor the ambient temperature is used to determine the temperature difference. The solar radiation is a substantial heat gain to the rooms.

3.2.1 Component: Ambient Temperature and Humidity

For the first three output signals (temperature, humidity and solar radiation) it is possible to choose between different ambient profiles. These profiles have their own different set of output data:

STATIC ambient conditions: for analysing static design conditions (worst case weather conditions). The available static profiles are:

- STATIC-S: summer design conditions from Table 1 - STATIC-W: winter design conditions from Table 1 - STATIC-user: user defined constant ambient conditions

REAL ambient conditions: for analysing ‘real world’ scenarios. The data is loaded from input tables. They are hourly average ambient conditions for different locations during a day (Figure 3-4). The available locations are:

Analysis of the impact of adjustments to the on board HVAC design by time domain simulations

Analysis of the impact of adjustments to the on board

HVAC design by time domain simulations

Master thesis

by

Analysis of the impact of adjustments to the on board

HVAC design by time domain simulations

CONFIDENTIAL

ABSTRACT

PREFACE

TABLE OF CONTENTS

LIST OF FIGURES

LIST OF TABLES

LIST OF SYMBOLS

LIST OF ABBREVIATIONS

1 INTRODUCTION

1.1 Background

1.2 Problem Definition

1.3 Research Objectives

1.4 Methodology and Materials

1.5 Project Limitations

2 HVAC SYSTEMS AND HVAC DESIGN

2.1 HVAC System

2.2 Main Components

2.3 HVAC Processes

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2.4 HVAC System AMELS 272

2.5 HVAC System Design Method

3 MODELLING OF THE HVAC SYSTEM

3.1 Overall HVAC System Model

3.2 Submodel: Ambient and Ship Conditions