key: cord-1023482-9stsen5w
authors: Afrasiabian, Ehsan; Douglas, Roy; Geron, Marco; Cunninghamb, Gareth
title: A numerical evaluation of a novel recovery fresh air heat pump concept for a generic electric bus
date: 2022-02-07
journal: Appl Therm Eng
DOI: 10.1016/j.applthermaleng.2022.118181
sha: ca52fd2a7fe8dd2ad175c2217ac2efc0b3937188
doc_id: 1023482
cord_uid: 9stsen5w

Since the outbreak of the worldwide COVID-19 pandemic, public transportation networks have faced unprecedented challenges and have looked for practical solutions to address the safety concerns. In the wintertime, when natural ventilation is not an option, implementing an all-fresh air heating policy could be a viable solution to lower the transmission of such infectious disease, by reducing the density of pathogens and exposure time. Even though this will expectedly increase the energy demand, it is particularly crucial when the people safety is the first priority to operate any public transportation systems. To tackle the transmission issue, in this paper two concepts of an all-fresh-air Heat Pump (HP), namely a Baseline (BHP) and a novel Recovery (RHP) concepts were investigated. These two HP’s concepts could provide better ventilation inside the bus, compared to a Conventional (CHP) concept where the cabin air is re-circulated. To address the energy consumption concerns, the RHP concept is proposed to improve the performance of the system by recovering part of the cabin waste heat, without using any additional heat exchangers. Three different coupled models of a generic single-deck cabin and a heat pump system for each concept were developed in the Simscape environment of MATLAB (R2020b). The performance of these HPs was investigated to evaluate how an all-fresh air policy could affect the performance of the system in the BHP, and the energy-saving potential of the RHP concepts. The performance of the system was studied under different ambient temperatures ranging from -5 to 5 °C, and for low and moderate occupancy levels. Results show that employing the RHP and BHP significantly improved the ventilation rate per person by at least 102% and at most 125%, compared to a CHP concept with 50% of re-circulated air. Moreover, adopting the RHP concept also reduced the power demand by at least 8% and at most 11%, compared to the BHP for the selected fan and blower flow rates.

Since the emergence of COVID-19 and the pandemic it has created, mass transportation including buses have faced unprecedented challenges of providing people with safe, reliable, and functional services. The main challenge here is linked with the valid concerns of the spread of COVID-19 when a large number of people travel in a limited space [1, 2] . After the early ambiguous period of facing the new threat, now scientists have a better understanding of its fatality and transmission potentials, especially when people are in close vicinity and/or closed spaces. In fact, COVID-19 and other respiratory pathogens are believed to be spread mainly through contact with contaminated surfaces and by infectious droplets , and plausibly aerosols . So mask wearing and social ( > 10 ) ( < 10 ) distancing (between 1 and 2 meters) are strongly advised to reduce the transmission risks. Even though the air born transmission is still a matter of discussion, many authors have identified it as a plausible transmission mechanism in closed spaces [3, 4, 5, 6] . Droplets and aerosols are generated by the infected person via breathing, talking, sneezing, and coughing. Droplets are heavier particles that deposit after some time due to their weight while aerosols can be dragged by the air currents, remain suspended for hours, and be inhaled by susceptible individuals [3] . Several researchers have suggested that ventilation is essential to deal with the airborne transmission in closed spaces and an appropriate ventilation rate should be adopted to dilute the pathogens concentration [7, 8, 9] . If the ventilation inside closed spaces such as public transport systems is poor, meaning low fresh air rates into the space and high air recirculation, the pathogens can build up and travel long distances, increasing the infection risks. To address this concern, different administrators and leading organizations have advised transport operators to avoid air re-circulation or increase the fresh air flow rate as much as possible to make sure there is always enough fresh air flowing through the area [10, 11, 12, 13, 14] .

On the other hand, in electric vehicles with limited energy storage capacity, heating and air conditioning of the cabin could be the main power consumer, especially in extreme ambient conditions, that could reduce the driving range by more than 50% [15, 16, 17] . In recent years, there has been growing interest in the application of Heat Pumps (HPs) in the automotive industry due to electrification of powertrain and the loss of waste heat from internal combustion engines. In this regard, HPs' higher Coefficient of Performance (COP) and lower energy consumption, comparing to electric heaters, makes them a feasible alternative for heating in electric buses. To minimize heating loads when using HPs, it is common practice to use re-circulated air with a proper mixing ratio with fresh air. Zhang et al. [18] found that for a passenger electric car with maximum return-air condition, the heating demand reduced by 46.4-62.1% (when the ambient temperature was -5 and -20°C, respectively) compared to the all-fresh air condition. Pan et al. [19] evaluated that in an electric vehicle, utilizing re-circulated air could reduce heating energy in a HP system by 33-57% and increase the driving range by 11-30%.

To date, several numerical and experimental studies have investigated the Heating, Ventilation, and Air Conditioning (HVAC) of passenger vehicles [20, 21, 22] ; however, there is much less research conducted specifically on buses [23, 24, 25] . Numerical studies have focused on the cabin's thermal conditions, as well as the performance of the HVAC systems, mainly through detailed CFD [21, 26] or lumped-parameter models [27, 28, 29, 30, 31, 22] . Mezrhab and Bouzidi [22] developed a numerical nodal model based on the finite difference method to evaluate the cabin's thermal conditions of a passenger car in summer time. They described the cabin as a network of solid and fluid nodes representing the compartment body and the air inside it, respectively. They investigated the impacts of solar radiation and the radiative features of the vehicle on the temperature inside the cabin. Marcos et al. [27] proposed a simplified thermal model to calculate the thermal load for the air conditioning of a passenger car without including any models for the HVAC system. Torregrosa-Jaime et al. [28] [36] studied the employment of an air-to-air heat exchanger in which the fresh outside air could exchange both latent and sensible heat with air coming from the cabin through a separating membrane. However, in the both papers [35, 36] , the authors suggested recovering heat from the ventilation air by using a specialized heat exchanger.

In this study, we introduce a recovery heat pump concept that would address both the passengers' safety and energy demand concerns in a single-deck electric bus. The importance and originality of this study is that it suggests recovering waste heat from the ventilation air without employing any additional heat exchangers. This could significantly improve the performance of the heat pump when an all-freshair policy is in place (COVID-mode). The novel idea consists of forcing the warm air in the cabin to pass through the evaporator (on the cold side) and consequently to increase the evaporator temperature and the available heat to be absorbed by the refrigerant. To investigate the performance enhancement of the new recovery system, we developed a model by coupling a heat pump and a cabin of a generic bus submodels. Three heating concepts were modeled, namely: a Conventional Heat Pump (CHP), a Baseline

Heat Pump (BHP), and a Recovery Heat Pump (RHP). In the CHP, air is mixed with 50% of fresh air and re-circulated inside the cabin to reduce the energy consumption of the HP unit. While in both the BHP and RHP, the all-fresh-air policy was implemented to eliminate air re-circulation inside the cabin and lower the pathogens airborne transmission risks. In the BHP, the evaporator was solely exposed to the outside cold air while in the RHP, the warm cabin air was mixed with the outside cold air and then flowed through the evaporator before being vented outside. The system operation, performance, power demand, and the fresh air provided per person in the CHP, BHP, and RHP concepts were investigated.

The findings make an important contribution to the field of public transport during pandemics, when the air re-circulation should be avoided, by demonstrating the RHP's potential both to improve the ventilation rate and reduce the energy consumption compared to the CHP and BHP concepts, respectively.

Three different models were developed for the CHP, RHP, and BHP concepts in the Simulink and MATLAB environment (2020b). The Thermal, Moist Air, and Two-Phase physical domains from Simscape toolbox were used to build up two sub-models for each case, namely a heating unit and the cabin. The former, is based on a vapor compression cycle for a heat pump and the latter is formed by a network of thermal elements to describe the principle heat loss/gain/storage within the cabin or through the bus body. These two sub-models are coupled through the evaporators and condensers, while they also exchange heat with the ambient environment.

The heating unit sub-model is composed of the main components required for a vapor compression cycle namely; an evaporator, accumulator, Compressor (Comp), condenser, Electronic Expansion Valve (EEV), fans, blowers, and sensors, as shown in Fig. 1 . In this study, R134a flows through these components as the refrigerant and exchanges heat with moist air through the condenser and evaporator.

The refrigerant absorbs heat at the evaporator after the outside cold air is drawn through it by pulling fans. The superheated vapor passes through an accumulator installed at the exit of the evaporator to make sure solely the superheated vapor goes to the compressor and to prevent any droplets from deteriorating the compressor's operation. The low-pressure and cold superheated refrigerant becomes highpressure/temperature after compression. At the condenser, heat is exchanged between the refrigerant and the fresh (or mixed in the RHP) air moved by the blowers. The sub-cooled vapor passes through the expansion valve to enter the evaporator again as a low-pressure/temperature two-phase refrigerant and to complete the cycle. Fig. 1 shows that in the CHP case, fresh air from outside is mixed with the recirculated air inside the cabin while the evaporator is exposed to 100% outside cold air. This is a common practice to reduce the waste heat through vents as well as the energy consumption. As Fig. 2 illustrates, in the BHP model, both the evaporator and condenser are fed with 100% fresh air in order to deal with pandemic policies. This will expectedly increase the energy demand from the heating unit but is particularly crucial when the risk of contagious diseases is high, and health concerns are the first priority and requirement to operate any public transportation systems. Therefore as seen in Fig. 3 , while both the waste heat at the evaporator. This architecture is noteworthy as it does not need major modifications such as redesigning the system or adding new components, thus can be applied to the systems that are already in-service with minor modifications. 

Here stands for the volumetric efficiency, is the refrigerant density at the compressor's suction v r,suc line, is the frequency (Hz), and is the compressor's displacement volume (per revolution). In comp this study, a variable speed compressor was modeled using a controlled mass flow rate source where the vapor is compressed through an isentropic process and the compressor work is calculated in terms of the isentropic work and the overall compressor efficiency by Eq.2:

The evaporator and condenser are sized according to the compressor capacity to efficiently exchange the required heat. In this study, the evaporator and condenser both are cross-flow fin and tubes heat exchangers and their specifications are listed in Table 1 . The EEV, installed after the condenser, functions as a solenoid valve to keep the pressure difference on the hot and cold sides of the HP when the compressor goes OFF, as well as throttling the refrigerant flow while the HP is in operation. This device is modeled using a variable local restriction block from the Two-Phase library. It opens every time the system goes ON and the opening is regulated to provide enough superheating degree at the suction line of the compressor. The controller sends the ON signal to the compressor, EEV, fans, and blowers when the cabin temperature falls below the lower temperature threshold ( ), and the OFF signal l = set -1 °C when it exceeds the upper one ( ). Both fans and blowers are controlled volumetric flow u = set +1 °s ources that operate in line with the other components. However, when the compressor goes OFF blowers keep working with reduced capacity to make sure there is a continuous flow of fresh air into the cabin.

In this study, to ensure that a fresh air supply is consistently flowing through the vehicle, the air flow rate of the blowers falls to 20% of their full rate when the control signal is OFF.

The cabin is modelled as a constant volume chamber connected with a thermal network that represents the heat transfer, storage, and generation within the cabin and with the ambient. Fig. 4 shows how different thermal elements of the bus cabin were connected to form the thermal network that is incorporated into the volume chamber. As illustrated by Fig. 1-3 , this chamber is also linked to the ambient where the moist air is blown into the cabin and vented into the ambient. The temperature of the cabin is governed by Eqs. 3-6: a p,a cab =m + sol + amb + met + vent + HP

Here, is the metabolic rate, is the body surface (Dubois Area [37] ), is height and is m D h w weight of an average adult person. In this study a seated person was assumed to have the following parameters; , and . In Eq.3, the main loads entail m = 60 (W/m 2 ), h = 1.7 (m) w = 70 (kg) [38] the heat transfer with ambient through the convective and conductive mechanisms , heat ( amb ) generated by the people on board , and heat vented outside . In this equation, stands ( met ) ( vent ) HP for the heat delivered from condenser into the cabin, and the effects of thermal masses to store the heat are represented by . In this study and for the nominal operational conditions of the HP systems, the m solar radiation was assumed negligible in the winter time. The distinctive thermal conductivity coefficients of the body and windows are assumed uniform, as well as the internal and external convective heat transfer coefficients, as listed in Table 1 . 

To verify the numerical model a series of simulations was carried out for the BHP case where amb , and (low occupancy level) The results were compared with a refrigeration toolset = -5 °C = 3 .

called CoolPack toolkit (V1.5). This toolkit includes EESCoolTools (as an Engineering Equation Solver tool in refrigeration area), Refrigeration Utility, and Dynamic analysis tools that are suitable for designing, sizing, and analysing refrigeration systems. Here the instantaneous sub-cooling and superheating degrees were obtained from the Simulink dynamic model, and imported as input into the Coolpack toolkit. As illustrated in Fig. 6 with above-mentioned assumptions, the current model can accurately predict the instantaneous power requirement of the HP system to run with the specifications listed in Table 1 , as well as the amount of heat delivered into the cabin. Moreover, the predictions of the refrigerant mass flow rate of such a system are in good agreement in both models. For the required power, delivered heat, and the refrigerant mass flow rate predicted by Simulink and CoolPack toolkit the maximum error is approximately 2%. 

In order to evaluate the performance of the CHP, BHP, and RHP concepts, it is essential to establish 

For given blower and fan air flow rates, the system performance parameters including the COP and the heating capacity vary by changing the compressor speed and EEV opening. This is mainly due to the changing in the refrigerant mass flow rate, super heating, and sub-cooling degrees. To determine the optimum operating point for both the compressor and the EEV, a series of simulations was conducted for a range of compressor speeds and the expansion valve opening, for each operational condition. Fig. 6 shows the system operation map created for the BHP case with and low amb = 0 °C Here the average power demand takes into account both the compressor, fans, and blowers power and reads as:

A similar contour map was created for each simulation to find the optimum operating point for the compressor RPM and the EEV opening where the required power was minimum, and the following constrains were satisfied:

 The sub-cooling degree at the exit of the condenser should be .

The refrigerant should leave the evaporator at a minimum superheating degree .

( sup < 5 °)

 Required warm-up time is fixed and equals . 30′ ± 30" 

In practice, the waste heat from cabin could be controlled by varying both the total airflow rate and the mixing ratio of fresh and re-circulated air into the cabin. The lower this ratio, the lower the energy waste due to the ventilation. However, this lowers the air quality inside the cabin and over time increases the CO2 concentration. In the RHP and BHP concepts where the all-fresh air policy is applied, and in the absence of air re-circulation, the concerns over the air quality inside the cabin are reduced. Therefore, the air flow rates into the cabin could be reduced to a bare minimum to decrease waste heat. However, the air flow rates through the condenser and evaporator should be high enough to provide the required sub-cooling and super-heating degrees on the refrigerant side. On top of that, for the RHP architecture the ratio between the blower and fan flow rates is expected to be an important parameter as it should affect the potential of the evaporator to recover the heat from the warm air before being vented. Recently, Afrasiabian et al. [39] studied the effects of implementing sliding set-point temperature on the comfort level inside the cabin and the energy consumption, in different ambient temperatures. They showed that when , the power demand would reduce by about 61% for amb = -5 °C setamb = 12

comparing to the case where . They also showed that reduced set-point °C setamb = 22 °C temperatures would still provide people with acceptable comfort conditions as long as their normal outdoor clothing level was not reduced when people onboard. Therefore, in the current study three ambient temperatures ranging from up to were investigated and the set-point temperature -5

+5 °C

was set based on a sliding scale as . In all cases the initial temperature of the cabin set = amb +12 °C is assumed , and the blowers operate with reduced rate of 20% of their maximum rate i = amb +5 °C when the HPs are OFF to provide continuous fresh air into the cabin. . As expected, the CHP case is the most efficient system in terms of amb = 0 °C energy consumption, and the BHP is the most energy intensive system. As this figure demonstrates, the RHP case could successfully reduce the energy consumption when an all-fresh air policy is implemented in a bus. As anticipated for all the cases, and due to the lower metabolic heat gain, the required power for the low occupancy level is higher compared to the medium level. For amb = 0 °C using a BHP system increases power demand by at least 20% and at most 51%. While adopting a RHP demands 15 -30% more power, both are compared with a CHP system. It is evident that by increasing the blower rate, the required power increases. It could be attributed to the higher air flow rates being vented outside thus increasing the waste heat from cabin. However, different concepts will have different sensitivity. For example, for low occupancy level and when the CHP, b = 1800 (m 3 .hr -1 )

RHP, and BHP require 6, 14, 22% more power compared to lower blower flow rate , respectively. Furthermore, Fig. 7 demonstrates that the RHP could operate as a ( b = 900 (m 3 .hr -1 )) moderate option between the CHP and BHP, where its potential to reduce the power demand improves as the ratio increases. The Power Saving Potential (PES) by the RHP system is defined as

As shown in Fig. 7 , by increasing from to the PES parameter increased from 4 to / 0.25 0.5 21% at medium occupancy level. These figures are respectively 5 and 18%, at the low occupancy level.

Hereafter and for the rest of this study, the operating blower and fan rates are considered and 1350

. At the selected flow rates, for using the BHP concept increases power 3600 (m 3 .hr -1 ) amb = 0 °C demand by 36% (and 32%) for moderate (and low) occupancy levels. While adopting the RHP system demands just 24% (and 20%) more power, when compared with the CHP. This comparison shows the potential savings when adopting a RHP instead of a BHP system that could reduce the energy consumption by 12%. modes, which are coloured in blue and red, respectively. As evident in the boosting mode, a RHP system operates with higher evaporator temperatures compared to a BHP, resulting in a lower power requirement. It is due to the fact that in a RHP concept, warmer air (mixture of the cabin and ambient air) flows through the evaporator, therefore more heat is available for the refrigerant to absorb in higher temperature and pressure. As a result of higher temperature of refrigerant its vapour density at the suction line and its mass flow rate through the compressor increase by 5% compared to the BHP system, as shown in Fig. 9 . Furthermore as Fig. 9 demonstrates and for the sake of a valid comparison, for all these three cases the super-heating and sub-cooling degrees were kept at the same values . (approximatly ± 1 °C) The instantaneous cabin temperature and relative humidity are shown in Fig. 10 for both the warm-up and boosting periods. As it is demonstrated, in all cases the cabin condition could be reached and maintained within the acceptable range successfully. However, the temperature and RH in the BHP concept shows a tendency to the right, indicating relatively longer operating time compared to the CHP and RHP systems. 

r,av = ∫ r /Δ , Fig. 11 depicts the normalized power demand over six hours and under three different ambient conditions. The required power depends on a number of parameters including the heating capacity, COP, operation time, compressor and fan speed, ventilation strategies, and the thermal comfort policy.

As this figure indicates, adopting a sliding set-point temperature would be beneficial in terms of the power demand for all cases, and also the system could operate with less divergence from its optimum and nominal conditions. Moreover as evident in Fig. 11 , for all cases when the number of on-board people increases the required power is reduced by at least 19% and at most 22%, as the heat generated by the people would contribute more to the heat balance inside the cabin. Employing a BHP system increases the averaged power demand by at least 32% and at most 39%, compared to the CHP case. As this figure demonstrates, the HP architecture plays a substantial role in its performance, and adopting the RHP instead of the BHP suggests remarkable PES at least 10% for and , and by at least 39% (and at most 49%) for the RHP, and at least 43% (and at most 58%) for the BHP concepts. In both cases, the waste heat is considerably higher than a CHP system. It is because more heat is needed to be provided by the RHP and BHP to make up the heating demand and keep the cabins condition within the acceptable range. As Fig. 11 shows, the required power for a BHP and RHP concept to run with low occupancy level, respectively drop by 5 and 2% in , compared with when = 5 (°)

. Likewise and as depicted by Fig. 12, for the same occupancy level the normalized = -5 (°) averaged delivered heat into the cabin by the BHP and RHP system is 9 and 6% lower for amb = 5 °C in comparison with when . This difference in the percentage of the additional power amb = -5 °C requirement and the delivered heat could be due to the lower COP of both concepts in higher ambient temperatures, as shown in Fig. 13 , where the COP of the system is defined as:

The COP of both the BHP and RHP systems is higher for and , in comparison 4% = 3 amb = -5 °C with condition. Moreover, in spite of the fact that the power demand in the RHP concept amb = 5 °C is lower than BHP, its delivered heat into the cabin is just lower. As can be seen in 8 -11%

3 - 6%   Fig. 13 , the RHP improves the COP remarkably comparing to the BHP and CHP. This could be associated with its higher evaporator temperature, as illustrated in Fig. 8 . The COP of the RHP system increases by at least 15% for in low occupancy level and at most 19 % for amb = 5 °C ( = 3) amb with medium occupancy level , compared to the CHP. These figures, are lower in = -5 °C ( = 16)

the BHP system and falls between 9-14%. Regarding the ventilation rates, the BHP and RHP stand out and suggest significant improvements comparing to the air-re-circulated system, the CHP. This is implied by Fig. 15 where fresh air flow rates into the cabin in terms of LPS parameter (Litres per Person per Second) are depicted. As it is shown the LPS for RHP is at least 102% and at most 108% higher than the CHP case. Likewise, a BHP offers at least 111% and at most 125% higher LPS comparing with the CHP case. Again, the higher LPS of the BHP could be associated with the higher operation time, compared to the RHP system. It is believed that in the absence of air re-circulation inside the cabin in both the RHP and BHP concepts, this excess amount of fresh air would improve the cabin air quality and provide the passengers with a safer environment regarding the spread of contagious diseases such as the COVID-19 virus.

Comparing Fig. 11-Fig. 14 , we see that the BHP and RHP units work not only more efficiently but also more intensely to be able to keep the cabin temperature around the determined set-point. In fact, in colder ambient conditions, the evaporator should function in lower temperatures/pressures to be able to absorb enough heat from the ambient (or the mixture of ambient and cabin air in the RHP case).

Therefore, the refrigerant vapour density at the suction side of the compressor drops as well as the refrigerant mass flow rate that results in lower heating capacities in lower ambient temperatures. In this situation the compressor speed should increase to make up the heating demand and accordingly the EEV opening needs to be adjusted to provide enough super-heating degree, which are shown in Fig. 16 and Fig. 17 respectively. As Fig. 16 and Fig. 17 indicate, in all the ambient temperatures, the compressor of the BHP system should operate with the highest speed with at least 37% and at most 52% higher RPM, and widest EEV opening which is at least and at most 46% more than the CHP. In the RHP case, these figures are 38% considerably lower than the BHP but still the system runs with (at least -at most) higher 25 -38% speed and wider EEV opening, compared to the CHP case. It is mainly due to the fact that 31 -36% in each ambient temperature and for their respective set-points, the condenser conditions in the BHP and RHP are similar even though the evaporator temperatures are different. In the BHP concept, the evaporator is exposed to cold air flow from outside; however, in the RHP a mixture of the cabin warm and the outside cold air passes through the evaporator that lets the refrigerant evaporates in higher temperatures/pressures, as illustrated in Fig. 8 . This leads to lower required work from the compressor in the RHP case than the BHP case. This is also implied by higher refrigerant mass flow rates in the RHP which is at least 32% and at most 41% higher than the CHP system, as shown in Fig. 18 . In the CHP, the compressor speed and the refrigerant mass flow rates are the lowest as less heating capacity is needed compared with both the BHP and RHP, as shown by Fig. 12 . The increase in the compressor speed along with the longer operation time could explain the higher power demand from the BHP, which is followed by the RHP, both compared with the CHP system, as reflected in Fig. 11. 

In this study, we investigated how using a novel recovery concept (RHP) would increase the efficiency of a heat pump in a generic single-deck bus cabin, when an all-fresh air policy was implemented. The proposed RHP concept here is based on recovering waste heat from the ventilation air without using any additional heat exchanger. In this concept the fresh air is delivered into the cabin and passes through the evaporator to improve both the cabin air quality and the system performance.

The main motivation of avoiding air re-circulation in such a system is due to the valid health concerns on its plausible negative effects on the spread of infectious diseases like the COVID-19 virus in public transportation systems. Three different coupled models of a generic bus cabin, and a heat pump unit were created where the latter one could be a baseline, recovery, or a conventional system. Two different occupancy levels were studies for the ambient temperature ranging from -5 to 5 , while a sliding cabin °C set-point temperature was employed. Accordingly, the results showed that:  The RHP system increased the delivered fresh air into the cabin by least 102% (111% in BHP) and at most 108% (125% in BHP), compared to the CHP system. This excess amount of fresh air along with no air re-circulation policy could provide a safer environment for passengers and is expected to mitigate the risks of contagious diseases spread.  Adopting 100% fresh air policy for the cabin would increase the waste heat as well. Thus, in the RHP system at least 39% (43% in the BHP) and at most 49% (58% in the BHP) more heat was delivered into the cabin to maintain its temperature around the set-point, compared to the CHP system.  Compared to the CHP concept, the COP increased by at least 15% and at most 19 % in the RHP concept. These figures fell between 9-14% for the BHP concept.  Employing the RHP concept could reduce the demanded power for heating the bus cabin by at least 8% and at most 11%, compared to the case where a BHP system is adopted. These figures mean that the RHP concept could suggest the Potential Energy Saving (PES) of at least 10% and at most 15%.

 The operation times of the RHP and BHP concepts were 1-5% (at least -at most) and 8-17% higher than the CHP, respectively. 

The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

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☐ The authors declare the following financial interests/personal relationships which may be considered as potential competing interests